
OopightW 



COPYRIGHT DEPOSIT. 



Steam Power Plant 
Engineering 



BY 

G. F. GEBHARDT 

PROFESSOR OF MECHANICAL ENGINEERING, ARMOUR 

INSTITUTE OF TECHNOLOGY 

CHICAGO, ILL. 



THIRD EDITION, REVISED AND ENLARGED 
FIRST THOUSAND 



NEW YORK 

JOHN WILEY & SONS 

London: CHAPMAN & HALL, Limited 

1910 



r 



Copyright, 1908, 1910, 

BY 
G. F. GEBHARDT 






d 



Stanhope lpress 

P. H. GILSON COMPANY 
BOSTON. U.S.A. 



CI.A2732G6 



PREFACE 



This book is the outcome of a series of lectures delivered to the 
Senior class of the Armour Institute of Technology, Chicago, 111. 
It is primarily intended as a text-book for engineering students, 
but, it is hoped, will also be of interest to practicing engineers. 

The field embraced by the title is a large one and it has been 
necessary to limit the treatment to essential elements. Much of the 
matter contained in the author's original notes, including that relat- 
ing to steam engine design, valve gears, steam boiler design, and the 
like, has therefore been omitted. The numerous references appear- 
ing throughout the text and the appended bibliographies, which 
have been carefully compiled, are depended upon to extend the 
scope of the work. The standard codes of the American Society of 
Mechanical Engineers for conducting engine and boiler trials are in 
frequent demand by engineers and have therefore been included as 
an appendix. 

Authorities have been freely consulted and extensive use made of 
current engineering literature, due acknowledgment being made by 
footnote or reference whenever possible. 

The matter included is representative of American practice and no 
effort has been made to include any other except in a few special 
cases. 

The author wishes to express his obligations to Prof. Raymond 
Burnham for many valuable suggestions and corrections, and to 
Mrs. Julia Beveridge, librarian at Armour Institute, for assistance 
in compiling references. 



iii 



PREFACE TO SECOND EDITION 



A number of additional changes have been made to bring this 
work into accord with more recent practice. All the typographical 
and other errors discovered in the first edition have been corrected. 



PREFACE TO THIRD EDITION 



All obsolete matter has been discarded, considerable new material 
has been added throughout the book, and many of the chapters have 
been entirely rewritten. 






IV 



CONTENTS 



Page 
CHAPTER I. — Elementary Steam Power Plants 1-13 

1. General 1 

2. Elementary Non-Condensing Plant 2 

3. Non-Condensing Plant. Exhaust Steam Heating 5 

4. Elementary Condensing Plant 7 

5. Condensing Plant with Full Complement of Heat-Saving Appliances . . 10 

CHAPTER II. — Fuels and Combustion 14-67 

6. General 14 

7. Classification of Fuels 14 

8. Solid Fuels 14 

9. Coal 15 

10. Anthracite 15 

11. Semi-Anthracite 16 

12. Semi-Bituminous 16 

13. Bituminous 16 

14. Lignite ■. 17 

15. Peat or Turf 18 

16. Wood, Straw, Sawdust, Bagasse, Tanbark 19 

17. Composition of Coal 24 

18. Combustion 24 

19. Temperature due to Combustion 27 

20. Air Required for Combustion 28 

21. Calorific Value of Coal 31 

22. Heat Losses in Burning Coal 32 

23. Loss in the Dry Chimney Gases 33 

24. Loss due to Incomplete Combustion 36 

25. Loss of Fuel through Grate 37 

26. Superheating the Moisture in the Air 37 

27. Loss due to Moisture in the Fuel 37 

28. Loss due to Presence of Hydrogen in the Fuel 38 

29. Loss due to Visible Smoke 38 

30. Radiation and Minor Losses 39 

31. Size of Coal 39 

32. Washed Coai 40 

33. Purchasing Coal • 42 

34. Powdered Coal 44 

35. Depreciation of Powdered-Coal Furnaces 44 

36. Storing Powdered Fuel 45 

37. Rate of Combustion with Powdered Coal 45 

38. Cost of Pulverizing Coal 45 

39. Efficiency of Powdered-Coal Furnaces 46 

v 



vi CONTENTS 

CHAPTER II — Continued P AGE 

40. Furnaces for Burning Powdered Coal 47 

41. Draft for Powdered Fuel 47 

42. Types of Powdered-Coal Burners 47 

43. Pinther Apparatus for Burning Powdered Coal 48 

44. Schwartzkopff Apparatus for Burning Powdered Coal 49 

45. Aero-Pulverizer Apparatus for Burning Powdered Coal 49 

46. Triumph Apparatus for Burning Powdered Coal 50 

47. Fuel Oil 51 

48. Chemical and Physical Properties of Fuel Oil 52 

49. Efficiency of Boilers with Fuel Oil 54 

50. Comparative Evaporative Economy of Oil and Coal 55 

51. Types of Oil Burners 55 

52. Furnaces for Burning Oil Fuel 59 

53. Air vs. Steam as an Atomizing Medium 60 

54. Oil Pressure 63 

55. Oil Storage and Transportation 64 

56. Conclusions of the U. S. Naval Liquid Fuel Board 65 

57. Gaseous Fuels 66 

CHAPTER III. — Boilers 66-123 

58. General. . 68 

59. Classification 68 

60. Vertical Tubular Boilers 68 

61. Fire-Box Boilers 71 

62. Scotch-Marine Boilers 72 

63. Robb-Mumford Boiler 73 

64. Horizontal Return Tubular Boiler 74 

65. Babcock & Wilcox Boiler 81 

66. Heine Boiler 82 

67. Wickes Vertical Water-Tube Boiler 82 

67a. Parker Boiler 86 

68. Stirling Boiler 87 

69. Unit of Evaporation 88 

69a. Heat Transmission 90 

70. Heating Surface 92 

71. Horse Power of a Boiler 93 

72. Grate Surface 95 

73. Boiler and Furnace Efficiency . 98 

74. Boiler Performances 99 

75. Effect of Capacity on Efficiency 104 

76. Thickness of Fire 109 

77. Influence of Initial Temperature on Efficiency . . Ill 

78. Cost of Boiler and Settings 112 

79. Selection of Type , 112 

80. Grates 114 

81. Rocking Grates 115 

82. Blow-Offs 116 

83. Dampers 118 

84. Water Gauge 119 



CONTENTS vii 

CHAPTER III — Continued Page 

85. Fusible Plugs . . 121 

86. Mechanical Tube Cleaners . . 121 

CHAPTER IV. — Smoke Prevention, Furnaces, Stokers ,124-151 

87. General 124 

88. Mechanical Stokers 125 

89. Chain Grates 126 

90. Step Grates, Front Feed 130 

91. Step Grates, Side Feed 136 

92. Underfeed Stokers 138 

93. Down-Draft Furnaces 139 

94. Sprinkling Furnaces 141 

95. Dutch Ovens , 141 

96. Twin-Fire Furnace 142 

96a. Chicago Settings 143 

97. Wooley Smokeless Furnace 149 

98. Kent's Wing-Wall Furnace 149 

99. Burke's Smokeless Furnace 151 

100. Admission of Air above Fire 151 

101. Cost of Stokers and Furnaces 151 

CHAPTER V. — Superheated Steam; Superheaters 152-180 

102. General. 152 

103. Economy of Superheat 153 

104. Limit of Superheat 154 

105. Specific Heat of Superheated Steam 155 

106. Types of Superheaters 162 

107. Babcock & Wilcox Superheater 163 

108. Stirling Superheater 164 

109. Foster Superheater 165 

110. Independently Fired Superheaters 166 

111. Materials for Superheaters 170 

112. Extent of Superheating Surface . . 170 

113. Performance of Superheaters 174 

113a. Properties of Superheated Steam 180 

CHAPTER VI. — Coal and Ash-Handling Apparatus .181-206 

114. General 181 

115. Coal Storage . . ; .... 181 

116. Coal Conveyors 183 

117. Hand Shoveling . 183 

118. Bucket Conveyors 184 

119. Belt Conveyors 192 

120. Elevating Tower, Hand-Car Distribution 193 

121. Overhead Storage, Bucket Hoist 195 

122. Elevating Tower, Cable-Car Distribution 196 

123. "Vacuum " Ash-Handling System .".'..■ 196 

124. Cost of Handling Coal and Ashes 201 

125. Coal Hoppers 202 

126. Coal Valves 205 



vm CONTENTS 

Page 
CHAPTER VII. — Chimneys 207-244 

127. Chimney Draft 207 

128. Chimney Formulas 212 

129. Height of Chimneys for Boilers Using Oil Fuel 218 

130. Classification of Chimneys 218 

131. Guyed Chimneys 219 

132. Self-Sustaining Steel Chimneys 219 

133. Thickness of Plates , 220 

134. Riveting 223 

135. Stability of Steel Chimneys 223 

136. Brick Chimneys 224 

137. Thickness of Walls 226 

138. Core and Lining 230 

139. Materials for Brick Chimneys 230 

140. Stability of Brick Chimneys ; 231 

141. Custodis Radial Brick Chimney 234 

142. Steel-Concrete Chimneys 234 

143. Breeching 240 

144. Chimney Foundations 240 

145. Chimney Efficiencies 241 

146. Cost of Chimneys 243 

CHAPTER VIII. — Mechanical Draft 245-266 

147. General 245 

148. Steam Jets 245 

149. Parsons Smokeless Furnace 248 

150. Heinrich Smokeless Furnace 249 

151. Fan Draft 249 

152. Theory of Fans 252 

153. Determination of the Size of Fan 258 

154. Chimney vs. Mechanical Draft 261 

155. Balanced Draft 264 

CHAPTER IX. — Steam Engines 267-326 

156. Introductory 267 

157. The Ideal Engine 267 

158. Thermal Efficiency of the Actual Engine 273 

159. Mechanical Efficiency 275 

160. Heat Losses in the Actual Engine 278 

161. Loss due to Moisture in the Steam at Admission 278 

162. Loss due to Leakage 279 

163. Loss due to Cylinder Condensation 279 

164. Loss due to Clearance Volume 281 

165. Loss due to Incomplete Expansion and Compression 282 

166. Loss due to Wire Drawing , 284 

167. Loss due to Friction 284 

168. Effect of Increasing Boiler Pressure 286 

169. Receive r-Reheaters 287 

170. Jackets 288 

171. Single and Double Acting Engines 290 

172. High and Low Speed Engines 290 



CONTENTS ix 

CHAPTER IX — Continued Page 

173-4. High-Speed Single-Valve Engines 291 

175. High-Speed Multi- Valve Engines 297 

176. Medium and Low Speed Engines 299 

177. Compound Engines „ 300 

178. Triple and Quadruple Engines 305 

179. Influence of Condensing 307 

180. Throttling vs. Automatic Cut-Off 310 

181. Influence of Superheat 313 

182. Binary Vapor Engines 321 

183. Cost of Engines 326 

CHAPTER X. — Steam Turbines 327-396 

184. Classification 327 

184a. General Elementary Theory 328 

185. De Laval Turbine 331 

186. Elementary Theory, De Laval Turbine .... 333 

187. Terry Turbine 345 

188. Kerr Turbine 346 

189. Curtis Turbine 350 

190. Elementary Theory, Curtis Turbine 358 

191. Hamilton-Holzworth Turbine 362 

192. Westinghouse-Parsons Turbine 365 

192a. Allis Chalmers Turbine 372 

193. Elementary Theory, Parsons Turbine 373 

194. Low and Mixed Pressure Turbines 376 

195. Advantages of the Steam Turbine 382 

196. Simplicity 382 

197. Economy of Space and Foundation 382 

198. Absence of Oil in Condensed Steam 384 

199. Regulation 384 

200. Overload Capacity 384 

201. Efficiency and Economy 384 

202. First Cost 389 

203. Cost of Operation 392 

204. Influence of Superheat 393 

205. Influence of High Vacua 395 

CHAPTER XI. — Condensers 397-469 

206. General 397 

207. Function of the Condenser 398 

208. Classification of Condensers 400 

209. Common Jet Condenser 401 

210. Condensing Water, Jet Condensers 404 

211. Effect of Aqueous Vapor upon the Degree of Vacuum 405 

212. Injection Orifice 407 

213. Volume of Condenser Chamber 408 

214. Injection and Discharge Pipes . . 408 

215. Siphon Condensers 408 

216. Size of Siphon Condensers 409 

217. Ejector Condensers 410 



x CONTENTS 

CHAPTER XI — Continued Page 

218. Barometric Condensers 411 

219. Water-Cooled Surface Condensers 416 

220. Cooling Water, Surface Condensers 420 

221. Extent of Water-Cooling Surface 421 

222. Dry-Air Surface Condensers 428 

223. Quantity of Air for Cooling (Dry-Air Condenser) 429 

224. Saturated-Air Surface Condensers 430 

225. Evaporative Surface Condensers 433 

226. Location and Arrangement of Condensers 433 

226a. Independent System . . . . 434 

227. Central Condensing Systems 439 

228. High- Vacuum Systems 441 

229. Power Consumption of Condenser Auxiliaries . . . 447 

230. Cost of Condensers 450 

231. Most Economical Vacuum . 451 

232. Choice of Condensers 452 

233. Water-Cooling Systems 453 

233a. Cooling Pond 454 

233b. Spray Fountain .' 455 

234. Cooling Towers . 456 

235. Parallel Comparison of Fan and Natural-Draft Cooling Towers 460 

236. Cooling-Tower Calculations 460 

236a. Hygrometry . . . . . 468 

237. Tests of Cooling Towers . 468 

CHAPTER XII. — Feed-Water Purifiers and Heaters 471-522 

238. General . . 471 

239. Chemical Purification 476 

240. Boiler Compounds 476 

241. Use of Kerosene and Petroleum Oils in Boiler Feed Water 477 

242. Use of Zinc in Boilers 478 

243. Methods of Introducing Compounds 478 

244. Weight of Compound Necessary 478 

245. Mechanical Purification 479 

246. Thermal Purification 479 

247. Purifying Plants 480 

248. Economy of Preheating Feed Water 484 

249. Classification of Feed-Water Heaters 485 

250. Open Heaters 486 

251. Open Heaters and Purifiers 489 

252. Temperatures in Open Heaters 489 

253. Pan Surface Required in Open Heaters 491 

254. Size of Shell, Open Heaters 491 

255. Classification of Closed Heaters 492 

256. Closed Heaters, Water-Tube : .' 493 

257. Closed Heaters, Steam-Tube 494 

258. Heating Surface, Closed Heaters 496 

259. Heat Transmission, Closed Heaters 497 

260. Open vs. Closed Heaters. 504 

261. Through Heaters 505 



CONTENTS xi 

CHAPTER XII — Continued Page 

262. Induced Heaters 506 

263. Live-Steam Heaters and Purifiers 507 

264. Economizers ! 508 

265. Value of Economizers 511 

266. Factors Determining Installation of Economizers 512 

267. Feed-Water Temperature due to Use of Economizers 512 

268. Choice of Feed-Water Heating Systems 516 

CHAPTER XIII. — Pumps 522-574 

269. Classification of Pumps 522 

270. Boiler-Feed Pumps, Direct-Acting Duplex 524 

271. Boiler-Feed Pumps, Direct- Acting, Steam-Actuated Gear 527 

272. Air and Vacuum Chambers 529 

273. Water Pistons and Plungers 530 

274. Performance of Piston Pumps 531 

275. Size of Boiler-Feed Pump 539 

276. Steam-Pump Governors „• 541 

277. Feed- Water Regulators, Steam Pumps 541 

278. Power Pumps 543 

279. Injectors 545 

280. Positive Injectors 547 

281. Automatic Injectors 547 

282. Performance of Injectors 548 

283. Injector vs. Steam Pump as a Boiler Feeder 548 

284. Air Pumps 552 

285. Dean Wet- Air Pump 552 

286. Size of Wet-Air Pumps for Jet Condensers 553 

287. Edwards Air Pump 555 

288. Mullan Valveless Air Pump 556 

289. Alberger Rotative Dry-Air Pump 557 

290. Size of Wet-Air Pumps for Surface Condensers 558 

291. Size of Dry- Air Pumps for Surface Condensers 558 

292. Centrifugal Pumps 560 

292a. Hot-Well Pumps 560 

293. Volute Centrifugal Pumps 561 

294. Turbine Centrifugal Pumps 561 

295. Performance of Centrifugal Pumps 563 

296. Rotary Pumps 567 

297. Circulating Pumps 572 

298. Air Lift 572 

CHAPTER XIV. — Separators, Traps, and Drains 575-605 

299. Live-Steam Separators , 575 

300. Classification of Separators 576 

301. Reverse-Current Steam Separators 577 

302. Centrifugal Steam Separators 578 

303. Baffle-Plate Steam Separators 579 

304. Mesh Steam Separators 580 

305. Location of Separators 580 

306. Exhaust-Steam Separators and Oil Eliminators 581 

307. Exhaust Heads 585 



xii CONTENTS 

CHAPTER XIV — Continued Page 

308. Drips 586 

309. Low-Pressure Drips 586 

310. Size of Pipes for Low-Pressure Drips 588 

311. High-Pressure Drips 588 

312. Classification of Traps 588 

313. Float Traps 589 

314. Bucket Traps 590 

315. Dump or Bowl Traps 591 

316. Expansion Traps 592 

317. Differential Traps 594 

318. Location of Traps 596 

319. Drips under Vacuum 597 

320. Drips under Alternate Pressure and Vacuum 599 

321. The Steam Loop 600 

322. The Holly Loop 602 

323. Returns Tank and Pump 602 

324. Office Building Drains 603 

CHAPTER XV. — Piping and Pipe Fittings 606-668 

325. General 606 

326. Drawings 606 

327. Material for Pipes and Fittings 606 

328. Size and Strength of Commercial Pipe 608 

329. Screwed Fittings 610 

330. Flanged Fittings 610 

331. Pipe Coverings 616 

332. Expansion due to Temperature Variation 618 

333. Pipe Supports and Anchors 621 

334. General Arrangement of High-Pressure Steam Piping 622 

335. Main Steam Headers 629 

336. Flow of Steam in Pipes 632 

337. Equation of Pipes 636 

338. Friction through Valves and Fittings ^ 639 

339. Exhaust Piping, Condensing Plants 642 

340. Exhaust Piping, Non-Condensing Plant, Webster Vacuum Heating 

System 642 

341. Exhaust Piping, Non-Condensing Plant, Paul Vacuum Heating System 643 

342. Automatic Temperature Control 646 

343. Feed-Water Piping 647 

344. Flow of Water through Orifices, Nozzles and Pipes 650 

345. Stop Valves 655 

346. Automatic Non-Return Valves 658 

347. Emergency Valves 658 

348. Check Valves 660 

349. Blow-off Cocks and Valves 661 

350. Safety Valves 663 

351. Back-Pressure and Atmospheric Relief Valves 665 

352. Reducing Valves; Pressure Regulators 666 

353. Foot Valves 668 



CONTENTS xiii 

Page 
CHAPTER XVI. — Lubricants and Lubrication 669-689 

354. General 669 

355. Vegetable Oils 669 

356. Animal Oils and Fats 669 

357. Mineral Oils ■ 670 

358. Solid Lubricants 671 

359. Greases 671 

360. Qualifications of Good Lubricants 671 

361. Identification of Oils 672 

362. Gravity 672 

363. Viscosity 673 

364. Flash Point 673 

365. Burning Point or Fire Test 674 

366. Acidity 674 

367. Cold Test 674 

368. Friction Test 674 

369. Atmospheric Surface Lubrication • 675 

370. Intermittent Feed 675 

371. Restricted Feed 675 

372. Oil Bath 675 

373. Oil Cups 677 

374. Telescopic Oiler 677 

375. Ring Oiler 678 

376. Centrifugal Oiler 678 

377. Pendulum Oiler 679 

378. "Splash " Oiling 679 

379. Gravity Oil Feed 680 

380. Low-Pressure Gravity System 680 

381. Compressed-Air Feed . 680 

382. Cylinder Lubrication 682 

383. Cylinder Cups 682 

384. Hydrostatic Lubricator 683 

385. Forced-Feed Cylinder Lubricator 684 

386. Siegrist System of Lubrication 685 

387. Oil Filters 687 

CHAPTER XVII. — Finance and Economics — Cost of Power 690-729 

388. Records 690 

389. Output 690 

390. Load Factor 691 

391. Cost of Operation 693 

392. Fixed Charges 693 

393. Interest 693 

394. Depreciation 694 

395. Maintenance 699 

396. Taxes and Insurance 699 

397. Operating Costs 699 

398. Labor, Attendance, Wages 699 

399. Fuel 700 

400. Oil, Waste, and Supplies 703 

401. Repairs and Maintenance 703 

402. Cost of Power 703 



xiv CONTENTS 

Page 
CHAPTER XVIII. — Testing and Measuring Instruments 730-749 

403. General ; 730 

404. Weighing Fuel 730 

405. Measurement of Water 730 

406. Steam Meters 734 

406a. Pressure Gauges 735 

407. Temperature Measurements 736 

408. Power Measurements 741 

409. Flue-Gas Analysis 741 

410. Measurement of Moisture in Steam 745 

411. Fuel Calorimeters 747 

411a. Hamler-Eddy Smoke Recorder 749 

CHAPTER XIX. — Typical Specifications 750-773 

412. Specifications for a Cross Compound Non-Condensing Engine 750 

413. Specifications for a Return Tubular Boiler 754 

414. Specifications for a Condenser Plant 758 

415. Specifications for a Piping System 760 

416. Government Specifications for Purchasing Coal 769 

CHAPTER XX. — A Typical Steam Turbine Station — Commonwealth 

Edison Company, Chicago 774-787 

CHAPTER XXI. — A Typical Isolated Station — West Albany Power 
Station of the New York Central Railroad Company, West 
Albany, N. Y 788-797 

APPENDIX A. — General Bibliography — Power Plant Engineering 

and Design 798-821 

APPENDIX B. — A. S. M. E. Rules for Conducting Boiler Trials, Code 

of 1899 822-845 

APPENDIX C. — A. S. M. E. Rules for Conducting Steam Engine 

Tests 846-872 

APPENDIX D. — Steam Tables 873-876 

APPENDIX E. — Equivalent Values of Mechanical and Electrical 

Units 877 

APPENDIX F. — Miscellaneous Conversion Tables 878 

APPENDIX G. — Rules for Firemen Using Illinois and Indiana Coal 

in Hand-Fired Furnaces 879-880 

APPENDIX H. — Mollier Diagram 881-885 



LIST OF TABLES 



Page 

1. Composition of Typical American Coals 23 

2. Data Relative to Elements Most Commonly Met with in Connection with 

Combustion 26 

3. Weight of Air per Pound of Combustible as Indicated by the Percentage of 

C0 2 in the Flue Gas 31 

4. Heat Carried away by the Dry Chimney Gases per Pound of Combustible.. 34 

5. Loss Due to Incomplete Combustion of Carbon to Carbon Monoxide 35 

6. Effect of Washing on Bituminous Coals . 41 

7. Comparative Tests of Babcock & Wilcox Boiler. Lump Coal vs. Powdered 

Coal 46 

8. Analyses of Typical American Fuel Oils 52 

9. Boiler Efficiencies, Fuel Oil 54 

10. Tests of Fuel-Oil Burners 62 

11. Characteristics of Gaseous Fuels 67 

12. Required Hourly Evaporation per Boiler Horse Power at Various Feed 

Temperatures and Steam Pressures 96 

13. Ratio of Heating Surface to Grate Surface in Recent Boiler Installations .... 98 

14. Examples of Steam Boiler Tests 100 

15. Cost of Evaporating Water, Results of Actual Tests 105 

16. Air Spaces and Thickness of Grate Bars 114 

17. Values of c p at Atmospheric Pressure by Various Authorities 157 

18. Average Yearly Expense for Repairs for Cast-iron Superheaters 171 

19. Difference in Heat Efficiency of Superheaters Installed in Flue and Sepa- 

rately Fired Superheaters 176 

20. Decrease in Temperature of Gases of Combustion due to Superheater 

Installed in Flue 177 

21. Increase in Heat Efficiency of the Boiler due to Superheater 178 

22. Comparative Boiler Tests, Saturated vs. Superheated Steam, at Spring 

Creek Pumping Station of the Brooklyn Waterworks 179 

23. Density and Weight of Air and Chimney Gas at Various Temperatures. . . . 209 

24. Theoretical Draft Pressures in Inches of Water for Various Chimney Tem- 

peratures 210 

25. Test of a 100-Foot Steel Chimney ..213 

26. Chimney Formulas » 215 

27. Size of Chimneys for Steam Boilers 216 

28. Approximate Weight and Cost of Guyed Steel Stacks 219 

29. Steel Stack Dimensions 222 

30. Dimensions of Steel Chimney Foundations 242 

31. Dimensions of Brick Factory Chimneys 244 

32-33. Test of Steam Jet Blowers 248 

34. Sizes of Forced-Draft Fans 261 

35. Sizes of Induced-Draft Fans 262 

xv 



xvi LIST OF TABLES 

Page 

36. Steam Engine Efficiencies 274 

37. Mechanical Efficiencies of Engines 276 

38. Distribution of Friction Losses in Engines 285 

39. Performance of High-Speed Engines 296 

40. Performance of Saturated- Steam Engines, Compound 306 

41. Effect of Condensing on Engine Economy 310 

42. Per Cent Moisture Evaporated by Throttling 312 

43. Performance of Superheated-Steam Engines 315 

44. Effect of Superheat on Simple Engines 316 

45. Effect of Superheat on Compound Engines 317 

46. Effect of Superheat on Triple-Expansion Engines 318 

47. Record Steam Engine Performance. Superheated Steam 321 

48. Performance of Steam Turbines 390 

49. Pressures of Aqueous Vapor, Regnault 399 

50. Ratio by Weight of Cooling Water to Steam Condensed 406 

51. Size of Siphon Condensers . . 409 

52. Square Feet Cooling Surface Necessary to Condense One Pound of Steam 

under Different Conditions 428 

53. Test of a Pennel Saturated- Air Condenser 432 

54. Test of a Cast-iron Evaporative Surface Condenser 434 

55. Power Consumption of Condenser Auxiliaries 449 

56. Most Economical Vacuum for Steam Turbines 452 

57. Most Economical Vacuum for Piston Engines 453 

58. Properties of Saturated Air 468 

59. Influence of Thickness of Scale on Heat Transmission 472 

60. Water and Boiler Scale Analyses 473 

61. Boiler Defects, Report of Hartford Steam Boiler Inspection and Insurance 

Company 474 

62. Percentage of Saving for each Degree of Increase in Temperature of Feed 

Water 485 

63. Feed-Water Temperatures, Open Heaters 490 

64. Extent of Heating Surface, Closed Heaters 498 

65. Mean Temperature Difference, Closed Heaters 499 

66. Heat Transmission, Closed Heaters 502 

67. Economizer Tests 515 

68. Pump Duties for Various Efficiencies and Steam Consumptions 537 

69. Maximum Height to which Pumps can Raise Water by Suction — Tem- 

perature Constant 538 

70. Maximum Height to which Pumps can Raise Water — Temperature 

Variable 539 

71. Range of Working Pressure — Metropolitan Injectors 550 

72. Commercial Sizes of Air Pumps for Condensers 558 

73. Data Pertaining to Single-Stage Centrifugal Pumps 566 

74. Data Pertaining to Multi-Stage Centrifugal Pumps 567 

75. Tests of Steam Separators 576 

76. Dimensions of Standard Wrought-Iron Pipes 611 

77. Comparative Costs of Different Types of Flanges 614 

78. Dimensions of Standard Flanges 614 ! 

79. Dimensions of Extra Heavy Flanges 615 

80. Loss of Heat from Bare Pipes in Still Air 615 

81. Experiments on Pipe Coverings , 617 



LIST OF TABLES xvii 

Page 

82. Coefficients of Expansion — Piping Materials 620 

83. Comparison of Formulas for the Flow of Steam in Pipes 634 

84. Comparison of Formulas for the Flow of Steam in Pipes 635 

85. Flow of Steam in Pipes, Babcock 637 

86. Flow of Steam in Pipes, Sickles 638 

87. Equation of Pipes 640 

88. Specific Gravity of Lubricating Oils 673 

89. Properties of Lubricating Oils 670, 676 

90. Approximate Useful Life of Various Portions of Steam Power Plant 

Equipments 694 

91. Rates of Depreciation 695 

92. Depreciation Percentages — Chicago Traction Valuation Commission 696 

93. Cost of Labor for Street Railway Plants 701 

94. Cost of Labor for Tall Office Buildings 702 

95. Operating Costs per Kilowatt Hour, Typical British Electric Light and 

Power Plants , 709 

96. Operating Costs per Kilowatt Hour, Typical United States Railway Plants 709 

97. Operating Costs per Kilowatt Hour, Average of all Stations, Boston Ele- 

vated • . 710 

98. Operating Costs (1907), First National Bank Building, Chicago 710 

99. Cost of One Horse Power per Year, Simple Engine, W. O. Webber 711 

100. Cost of One Horse Power per Year, Compound Engine, W. O. Webber. . . 711 

101. Cost of One Horse Power per Year, H. von Schon 712 

102. Cost of Electrical Power per Year, W. M. Wilson 713 

103. Cost of Electrical Power per Year, R. C. Carpenter 715 

104. Cost of Electrical Power per Year, Oil Fuel, C. C. Moore & Co 716 

105-7. Cost of Power, Typical Isolated Stations 720, 722 

108. Cost of Power, First National Bank Building, Chicago 724 

109. Temperature Ranges of Thermometers in General Use 739 

Appendix D . — Steam Tables 873 

Additional Tables — Third Edition. 

00. Physical and Chemical Properties of Woods, Straw and Tanbark 19 

0. Heat Values of Bagasse and Variation with Degree of Extraction 20 

2a. Ratio of Total Air Supplied to that Theoretically Required for Various 

Analyses of Flue Gases 30 

14a. Principal Data and Results, Boiler Unit No. 10, Fisk Street Station, Com- 
monwealth Edison Company 102 

15a. Pounds of Water Evaporated per Hour from and at 212° F. per Pound of 

Fuel 108 

17a. Mean Specific Heat of Superheated Steam 159 

17b. Specific Volume of Superheated Steam 160 

94a. Distribution of Maintenance and Operating Costs in Large Stations. .... 707 

107a. Yearly Operating Costs in Four" Typical Central Stations 723 

107b. Power Costs — Steam Electric Central Stations 723 



ILLUSTRATIONS. 



Chapter I 

Elementary Steam Power Plants. 
Fig. 

1. Elementary Non-Condensing Plant 

2. Elementary Non-Condensing Plant with Heating System. 

3. Simple Condensing Plant. 

4. Condensing Plant with Full Complement of Heat-Saving Appliances. 
4a. Sectional Elevation of the Myers Furnace for Burning Bagasse. 

4b. End and Side Sectional Elevation of the Myers Furnace for Burning Tanbark, 

Chapter II. — Fuels and Combustion. 

5. Relation of Gas Composition in Combustion Chamber to Temperature. 

6. Influence of Size of Coal on Boiler Capacity and Efficiency. 

7. Influence of Ash on Fuel Value of Dry Coal. 

8. Pinther Coal-dust Feeder. 

9. Schwartzkopff Coal-dust Feeder. 

10. Aero-Pulverizer Coal-dust Feeder. 

11. Triumph Coal-dust Feeder. 

12. Korting Fuel Oil Burner. 

13. Booth Fuel Oil Burner. 

14. Hammel Fuel Oil Burner. 

15. Branch Fuel Oil Burner. 

16. Kirkwood Fuel Oil Burner. 

17. Williams Fuel Oil Burner. 

18. Warren Fuel Oil Burner. 

19. Furnace for Burning Fuel Oil, Front Feed. 

20. Furnace for Burning Fuel Oil, Rear Feed. 

21. International Gas and Fuel Company's Fuel Oil System. 

22. Hydraulic Oil Storage Company's Fuel Oil System. 

Chapter III. — Boilers. 

23. Vertical Tubular Boiler with Submerged Tube Sheet. 

24. Manning Vertical Fire-Tube Boiler. 

25. Typical Fire-box Boiler, — Stationary Type. 

26. Stationary Scotch-Marine Boiler. 

27. Robb-Mumford Boiler. 

28. Return Tubular Boiler Setting, — Extended Front. 

29. Return Tubular Boiler Setting, — Flush Front. 

30. Return Tubular Boiler Setting, — Steel Beam Suspension. 

31. Boiler Setting, "Wood" Mill of the American Woolen Company, Lawrence, 

Mass. 

32. Furnace Arch Bars. 

33. Back Connection made with Cast-iron Plate. 

34. Babcock and Wilcox Boiler and Setting. 



xx ILLUSTRATIONS 

Fig. 

35. Details of Header, — Babcock and Wilcox Boiler. 

36. Front Section, — Babcock and Wilcox Boiler. 

37. Heine Boiler and Setting. 

38. Wickes Vertical Water-Tube Boiler. 
38a. 1200-H.P. Parker Boiler. 

38b. Boiler Room Area for Various Types of Boilers. 

39. Stirling Boiler and Setting. 

39a. Heat Transmission Through Boiler Plate. 

39b. Influence of Draft on Capacity, Torpedo Boat "Biddle." 

40. Influence of Draft on the Efficiency and Capacity of a 350-Horse-power 

Babcock and Wilcox Boiler with Chain Grate. 

41. Effect of Rate of Driving on Economy of a 150-Horse-power Stirling Boiler, 

Hand Fired. 
41a. Relation Between Efficiency and Capacity, 500-Horse-power Babcock and 

Wilcox Boiler. 
41b. Effect of Rate of Driving on Efficiency of a 600-Horse-power Babcock and 

Wilcox Boiler. 
41c. Influence of Draft on the Capacity of a 600-Horse-power Babcock and Wilcox 

Boiler. 

42. Effect of Thickness of Fire on the Capacity and Efficiency of a 350-Horse- 

power Stirling Boiler, equipped with Chain Grate. 

43. Effect of Thickness of Fire on the Capacity and Efficiency of a 150-Horse- 

power Water-Tube Boiler. 

44. Effect of Thickness of Fire on the Capacity and Efficiency of a 500-Horse- 

power Babcock and Wilcox Boiler. 

45. Types of Grate Bars. 

46. A Typical Rocking Grate. 

47. Horizontal Blow-off Connection to Head. 

48. Vertical Blow-off Connection to Shell. 

49. Blow-off Connection with Circulating Pipe. 

50. Blow-off Tank and Connections. 

51. Surface Blow-off. 

52. Buckeye Skimmer. 

53. Kitts Hydraulic Damper Regulator. 

54. Tilden Steam Actuated Damper Regulator. 

55. Simple Water Column. 

56. Water Gauge with Self-closing Valve. 

57. Combined Water Column and High and Low Water Alarm. 

58. Types of Fusible Plugs. 

59. Mechanical Tube Cleaner, — Hammer Type. 

60. Mechanical Tube Cleaner, — Turbine Type. 

Chapter IV. — Smoke Prevention, Furnaces, Stokers. 

61. Green Chain Grate. 

62. Babcock and Wilcox Boiler, Chain Grate, Ordinary Setting. 

63. Babcock and Wilcox Boiler, Chain Grate, Fire-tile Roof. 

64. Section of Fire Tile. 

65. Section of Fire Tile. 

66. Application of "Economy" Fire Files to Stirling Boiler. 

67. Method of Anchoring "Economy" Fire Tiles to Tubes. 
67a. Chain Grate Fired from Rear End of Setting. 



ILLUSTRATIONS xxi 

Fig. 

67b. Smokeless Setting, Chain Grate and Babcock and Wilcox Boiler. 

68. Details of Roney Stoker. 
68a. Double Stoker Setting. 

69. Details of Wilkinson Stoker. 

70. Murphy Furnace, Front Section. 

71. Murphy Furnace, Side Section. 

72. Jones Underfeed Stoker. 

73. American Underfeed Stoker. 

74. Hawley Down-Draft Furnace. 

75. Plain Dutch Oven. 

76. "Twin Fire Arch," Applied to a Return-Tubular Boiler. 

76 a, b, c. Chicago Setting, Hand-Fired Return-Tubular Boiler. 

77. Wooley Smokeless Furnace. 

78. Kent's Wing-Wall Furnace. 

79. Burke's Smokeless Furnace, Front Section. 

80. Burke's Smokeless Furnace, Side Section. 

81. Split Bridge Wall. 

Chapter V. — Superheated Steam; Superheaters. 

82. Specific Heat of Superheated Steam, Knoblauch and Linde. 

83. Specific Heat of Superheated Steam, A. R. Dodge. 

84. Specific Heat of Superheated Steam, C. E. Burgeon. 

85. Specific Heat of Superheated Steam, Thomas. 

86. Babcock and Wilcox Superheater. 

87. Stirling Superheater. 

88. Details of Stirling Superheater. 

89. Foster Superheater in Babcock and Wilcox Boiler. 

90. Schmidt Independently Fired Superheater. 

91. Foster Independently Fired Superheater. 

92. Schmidt System of Combined Superheater, Economizer and Feed-Water 

Heater. 
92a. Relation between Gas Temperature, Heating Surface passed over and Amount 
of Steam Generated. 

93. Relation of Superheat to Total Output of Boiler. 

94. Relation of Output of Superheater to Boiler Output. 

95. Relation of Superheat to Output of Superheater. 

Chapter VI. — Coal and Ash Handling Apparatus. 

96. Link-Belt Coal-Handling Apparatus. 

97. Typical Coal and Ash Handling Equipment. 

98. Steel Cable Company's Coal-Handling Apparatus. 

99. Coal and Ash Handling System of S. S. Elevated Railway. 

100. Crusher and Conveyor of S. S. Elevated Railway. 

101. Driving Mechanism of Hunt Conveyor. 

102. Hunt Coal Conveyor System at Baltimore, Md. 

103. Bucket and Screw Conveyor at Commercial National Bank Building, Chicago, 

Illinois. 

104. Guide Pulleys, Robins Belt Conveyor. 

105. Coal and Ash Handling System of Aurora and Elgin Railway. 

106. Coa] and Ash Handling System of Cincinnati Traction Company. 



xxii ILLUSTRATIONS 

Fig. 

107. Coal and Ash Handling System of Detroit Edison Company. 

108. Vacuum Ash Handling System. 

108a. Vacuum Ash Handling System at the Armour Glue Works. 
108b. Coal and Ash Handling System, Norfolk Traction Company. 

109. Stationary Coal-Weighing Hoppers. 

110. Traveling Coal Hoppers. 

111. Common Slide Coal Valve. 

112. Simplex Coal Valve. 

113. Duplex Coal Valve. 

114. Flap Coal Valve. 

115. Seaton Coal Valve. 

Chapter VII. — Chimneys. 

116. Relation between Draft and Rates of Combustion. 

117. Steel Chimney of S. S. Elevated Railway Power House, Chicago. 

118. Stability of Steel Chimneys. 

119. Custodis Radial Brick Chimney. 
119a. Custodis Radial Perforated Brick. 

120. Circular Brick Chimney at Armour Institute of Technology. 

121. Design of Brick Chimneys, Thickness of Walls. 

122. Design of Brick Chimneys, Stability. 

123. Weber Steel-Concrete Chimney. 

124. Weber Steel-Concrete Chimney. 

Chapter VIII. — Mechanical Draft. 

125. Ring Steam Jet. 

126. Bloomsburg Jet. 

127. McClaves Argand Blower. 

128. Hollow Bridge Wall and Steam Jet. 

129. Parsons Smokeless Furnace. 

130. Heinrich Smokeless Furnace. (Sectional Elevation.) 

131. Heinrich Smokeless Furnace. (Sectional Plan.) 

132. Typical Forced-Draft System. 

133. Typical Induced-Draft System. 

134. Pitot Tubes; Orifice Closed. 

135. Pitot Tubes; Orifice Wide Open. 

136. Pitot Tubes; Orifice Partly Closed. 

137. Performance of Steel Plate Fans. 

138. Performance of Pressure Blower, Speed Constant. 

139. Performance of Pressure Blower, Speed Variable. 

140. Comparative Costs of Chimneys and Mechanical Draft. 

141. Influence of Rate of Combustion on Air Supply; Forced Draft. 

142. Balanced Draft System. 

Chapter IX. — Steam Engines. 
142a. Rankine Cycle. 

142b. Side Elevation, Typical Corliss Engine. 
142c. Plan View, Typical Corliss Engine. 
142d. A Modern Piston Engine Plant. 

143. Mechanical Efficiencies of Engine and Generator. 

144. Status of the Steam Engine. 

145. Condensation and Leakage Losses in Simple Engines. 



ILLUSTRATIONS xxiii 

Fig. 

145a. Influence of Increasing Back Pressure. 

146. Typical Curves of Steam Engine Friction. 

147. Influence of Increasing Initial Pressure. 

148. Typical Economy Curves — High-Speed Engines. 
148a. Assembly of Valve Gear, Typical Corliss Engine. 
148b. Section through Cylinder, Typical Corliss Engine. 

148c. Assembly of Governor and Link Mechanism, Corliss Engine. 

149. Test of Reeves Simple Engine; Condensing vs. Non-Condensing. 

150. Typical Economy Curves of Single- Valve vs. Four- Valve High-Speed Engines. 
150a. 3500-K.W. Vertical Cross Compound Corliss Engine. 

150b. 7500-K.W. Vertical Horizontal Cross Compound Corliss Engine. 

151. Effect of Compounding on High-Speed Non-Condensing Engines. 

152. Performance of Corliss Compound; Condensing vs. Non-Condensing. 

153. Performance of a 5500-Horse-power Engine. 

154. Increase in Power Due to Vacuum. 

155. Increase in Power Due to Vacuum. 

156. Performance of a 5500-Horse-power Engine. 

157. Indicator Cards — High-Speed Throttling Engines. 

158. Indicator Cards — High-Speed Automatic Engines. 

159. Effect of Superheat on Steam Consumption. 

159a. 3000-Horse-power Sulzer Engine for Highly Superheated Steam. 
159b. Fleming-Harrisburg Four- Valve Tandem Compound. 

160. Effect of Superheat on Steam Compounds. 

161. Influence of Superheat on Economy. 

162. Diagrammatic Arrangement, Binary- Vapor Engine. 

163. Cost of Simple High-Speed Engines. 

164. Cost of High-Speed Compound Engines. 

165. Cost of Simple and Low-Speed Compound Engines. 

Chapter X. — Steam Turbine. 

166. Horizontal Section of De Laval Turbine. 

167. Details of Blades of De Laval Turbine. 

168. Details of Nozzle of De Laval Turbine. 

169. Details of Governor of De Laval Turbine. 

170. Theoretically Proportional Expanding Nozzle. 

171. Theoretical Performance of a Divergent Nozzle. 

172. Characteristic Performance of a Divergent Nozzle. 
172a. Velocity Diagram, Ideal Impulse Turbine. 

172b.. Velocity Diagram, as Modified by Friction Losses. 

173. Section through Terry Turbine. 

174. Arrangement of Buckets and Reversing Chambers, Terry Turbine. 

175. Longitudinal Section through Kerr Turbine. 

176. Sectional End Elevation, Kerr Turbine. 

177. Details of Governor, Kerr Turbine. 

178. Four-Stage Vertical Curtis Turbo-Generator. 

179. 3500-K.W. Horizontal Curtis Turbine. 

180. Arrangement of Nozzles and Blades, Curtis Turbine. 

181. Section through Curtis Governor. 

182. Mechanical Valve Gear, Curtis Turbine. 

183. Hydraulic Valve Gear, Curtis Turbine. 

183a. Steam Belt Area in Five-Stage Curtis Turbine. 



xxiv ILLUSTRATIONS 

Fig. 

184. Velocity Diagram, Curtis Turbine. 

185. Section through Hamilton-Holzworth Turbine. 

186. Details of Vanes, Hamilton-Holzworth Turbine. 

187. Details of Bearings, Hamilton-Holzworth Turbine 

188. Details of Governor, Hamilton-Holzworth Turbine. 

189. Section through Westinghouse-Parsons Standard Turbine. 

190. Flow of Steam in Parsons Turbine. 

191. Details of Governor, Westinghouse-Parsons Turbine. 

192. Indicator Cards, Westinghouse-Parsons Turbine. 

193. By-Pass Valve, Westinghouse-Parsons Turbine. 

193a. Method of Fastening Blades, Westinghouse-Parsons Turbine. 
193b. High-Pressure Double-Flow, Westinghouse-Parsons Turbine. 
193c. Allis-Chalmers Steam Turbine. 

194. Velocity Diagram, Multi-Stage Reaction Turbine. 

194a. Low-Pressure Turbine, 59th St. Station, Interborough Rapid Transit Company. 

194b. Low-Pressure Double-Flow Westinghouse-Parsons Turbine. 

194c. Performance of 7500-K.W. Engine at 59th St. Station. 

194d. Comparison of Economy Curves, Combined Engine and Turbine. 

195. Rateau Low-Pressure Steam Turbine at South Chicago. 

196. Rateau Regenerator Accumulator. 

196a. Typical Double-Deck Turbine Installation. 

197. Curve of Performance of Rateau Low-Pressure Turbine. 

198. Comparative Floor Space, Engines vs. Turbines. 

198a. Typical Performance of 90-Horse-power Terry Turbine. 
198b. Typical Performance 9000-K.W. Curtis Turbine. 
198c. Typical Performance Small Non-Condensing Turbines. 
198d. Typical Correction Curves, 125-K. W. Turbines. 

199. Reciprocating Engine vs. Turbine Economy. 

200. Effect of Superheat on Economy. 

201. Effect of Vacuum on Economy, Westinghouse-Parsons Turbine. 
201a. Effect of a Vacuum on Economy. 

202. Effect of Vacuum and Superheat on Economy. 

Chapter XL — Condensers. 

203. Worthington Jet Condenser. 

204. Blake Jet Condenser. 

205. Baragwanath Siphon Condenser. 

206. Schutte Ejector Condenser. 

207. Piping for Schutte Condenser. 

208. Weiss Counter-Current Condenser. 

209. Alberger Barometric Condenser. 

210. Worthington Barometric Condenser. 

211. Tomlinson Type B Barometric Condenser. 

211a. Centrifugal Pump Applied to Tail Pipe of a Barometric Condenser. 

212. Baragwanath Surface Condenser. 

213. Wheeler Surface Condenser and Pumps. 

214. Wheeler Multi-flow Surface Condenser. 

215. Weighton Multi-flow Surface Condenser. 

216. Relation between Hot-well Temperature and Vacuum in Surface Condensers. 
216a. Application of Weighton Dry-Tube Surface Condenser to Vertical Marine 

Ensrine. 









ILLUSTRATIONS xxv 

Fig. 

216b. Heat Transfer in Condenser Tubes, Steam to Water. 

216c, d. Heat Transfer in Condenser Tubes, Steam to Air. 

217. Pennel Saturated-Air Surface Condenser. 

218. Pennel Flask Type Atmospheric Condenser. 

219. Jet Condenser located below Engine-Room Floor. 

220. Surface Condenser located below Engine-Room Floor. 

221. Surface Condenser Connected with Pumping Engine. 

222. Jet Condenser located above Engine-Room Floor. 

223. Typical Arrangement, Westinghouse-Leblanc Condenser and Curtis Turbine, 

224. Elevation of Condenser Piping, Des Moines City Railroad Power House. 

225. Plan of the Condenser Piping, Des Moines City Railroad Power House. 

226. Plan of Condenser Piping, Northwestern Elevated Railroad Power House, 

Chicago. 
226a. Condenser Installation, Quincy Point Power Plant. 

227. Worthington High-Vacuum System. 

228. Wheeler High- Vacuum System. 

229. Parsons Vacuum Augmenter. 

229a. Westinghouse-Leblanc High Vacuum Multi-Jet Condenser. 
229b. Tomlinson Type C High Vacuum Jet Condener. 
229c. Korting Multi-Jet Condenser. 

230. Power Consumption of Auxiliaries, Parsons Turbine. 

231. Power Consumption of Auxiliaries, Curtis Turbine. 

232. Relative Cost of High- Vacuum Condensing Systems. 

233. Performance of Spray Fountain. 

234. Barnard-Wheeler Cooling Tower. 

235. Worthington Cooling Tower. 

236. Alberger Cooling Tower Installation. 



Chapter XII. — Feed-Water Heaters. 

237. Scaife System for Feed-Water Purification. 

238. We-Fu-Go System for Feed-Water Purification. 
238a. Anderson System for Feed-Water Purification. 

239. Cochrane Feed-Water Heater and Receiver. 

240. Webster Star Vacuum Heater. 

241. Hoppes Horizontal Heater and Purifier. 

242. Goubert Single-flow Heater. 

243. Expansion Joint, Goubert Heater. 

244. Wainwright Multi-flow Closed Heater. 

245. Coil Heater. 

246. Otis Steam-Tube Feed-Water Heater. 

247. Baragwanath Steam-Jacketed Heater. 

248. Heat Transmission in Feed-Water Heater Tubes. 

249. Open Heater Connected as a Through Heater 

250. Through Heater with By-Pass. 

251. Open Induced Heater, Non-Condensing Plant. 

252. Closed Induced Heater, Condensing Plant. 

253. Hoppes Live-Steam Heater and Purifier. 

254. Installation of a Live-Steam Purifier. 

255. Typical Installation of Primary and Secondary Heater. 



xxvi ILLUSTRATIONS 

Fig. 

256. Green Economizer. 

257. Economizer Installation at Weehawken, New Jersey. 

258. Heat Transmission, Economizers. 

Chapter XIII. — Pumps. 

259. Duplex Direct-Acting Boiler-Feed Pump. 

260. Section through Duplex Boiler-Feed Pump. 

261. Method of Obtaining Lost Motion, — Duplex Valve Gear. 

262. Method of Obtaining Lost Motion, — Duplex Valve Gear. 

263. Position of Valve and Piston at the Beginning of Stroke. 

264. Position of Valve and Piston at the End of Stroke. 

265. Pump Disk Valve. 

266. Section through a Compound Duplex Pump. 

267. Section through a Simplex Pump with Steam-Actuated Gear. 

268. Forms of Vacuum Chambers. 

269. Different Arrangements of Vacuum Chambers. 

270. Types of Water Pistons. 

271. Plunger with Metal Packing Ring. 

272. Plunger with Hydraulic Packing. 

273. Horizontal Fly-Wheel Pump with Outsicre Packed Plunger. 

274. Performance of Direct-Acting Pressure Pumps. 

275. Performance of Boiler-Feed Pump at the Armour Institute of Technology. 

276. Fisher Pump Governor. 

277. Kitts Feed- Water Regulator. 

278. Rowe Feed-Water Regulator. 

279. Triplex Pump. 

279a. Performance of Triplex Pump, Direct Connected. 

280. Performance of Triplex Pump, Geared. 

281. Elementary Form of Ejector. 

282. The Hancock Inspirator. 

283. The Penberthy Automatic Injector. 

284a. Performance of a Desmond Automatic Injector with Varying Initial Pressure. 
284b. Performance of a Desmond Automatic Injector with Varying Suction 

Temperature. 
284c. Performance of a Desmond Automatic Injector with Varying Discharge 

Pressure. 

285. Dean Jet Condenser Air Pump. 

286. Edwards Air Pump. 

287. Mullan Valveless Air Pump. 

288. Hewes and Phillips Air Pump. 

289. Alberger Dry- Air Pump. 

290. Air Pump Indicator Diagram. 

291. Types of Impellers, Centrifugal Pumps. 

292. A Typical Centrifugal Pump. 

293. Direction of Water from Impeller of Volute Pumps without Diffusion Vanes. 

294. Effect of Diffusion Vanes on the Direction of Water. 

295. Three-Stage Lea-Degan Turbine Pump. 

296. Six-Stage Rateau Turbine Pump. 

297. Test of Centrifugal Pump at Armour Institute. 

298. Centrifugal Pump Characteristic for Boiler-Feed Pumps. 



ILLUSTRATIONS xxvii 

Fig. 

299. Centrifugal Pump Characteristic for Dry-Dock Service. 

300. Centrifugal Pump Characteristic for Waterworks Service. 

301. Performance of a 6-inch Worthington Conoidal Centrifugal Pump. 

302. Performance of a Single-Stage De Laval Volute Pump. 

303. Performance of a Two-Stage Turbine Pump. 

304. Performance of a Two-Stage De Laval Centrifugal Pump. 

305. Two-Lobe Cycloidal Rotary Pump. 

306. Rotary Pump with Movable Butment. 

307. Performance of a Rotary Pump. 

308. High-Duty Circulating Pump, New York Rapid Transit Company. 

309. Pulsometer. 

310. Air Lift. 

Chapter XIV. — Separators, Traps, Drains. 

311. Hoppes Live-Steam Separator. 

312. Stratton Live-Steam Separator. 

313. Keystone Live-Steam Separator. 

314. Bundy Live-Steam Separator. 

315. Austin Live-Steam Separator. 

316. Direct Live-Steam Separator. 

317. Baum Oil Separator. 

318. Loew Grease Extractor. 

319. Typical Exhaust Head. 

320. Closed Heater as a Blow-off Tank. 

321. Piping Drips to Exhaust Pipe. 

322. McDaniel Float Trap. 

323. Acme Bucket Trap. 

324. Bundy Tilting Trap. 

325. Columbia Expansion Trap. 

326. Geipel Expansion Trap. 

327. Dunham Expansion Trap. 

328. Heintz Expansion Trap. 

329. Flinn Differential Trap. 

330. Simple Siphon Trap. 

331. Location of Return Trap. 

332-3. Drainage for Jackets and Receivers of Triple Expansion Pumping Engines. 

334. Gravity Drainage; Vacuum Heater. 

335. Method of Draining Heater under Vacuum. 

336. Method of Draining Receivers under Alternate Pressure and Vacuum 

337. Steam Loop. 

338. Holly Loop. 

339. Section through Holly Receiver. 

340. Returns Tank and Pump. 

341. Shone Ejector. 

Chapter XV. — Piping and Pipe Fittings. 

342. United States Standard Pipe Thread. 

343. Types of Pipe Flanges. 

344. Efficiencies of Various Pipe Coverings. 

345. Pipe Bends. 

346. Double-Swing Screwed Fittings for Expansion. 



xxviii ILLUSTRATIONS 

Fig. 

347. Slip Expansion Joint. 

348. Typical Pipe Hanger. 

349. Typical Wall Bracket with Roll Binder. 
349a. A Typical Floor Stand. 

350. Typical Pipe Anchor. 

351. Arrangement of Steam Piping, Princeton University Power Plant. 

352. Typical "Duplicate" Header System. 

353. Typical "Loop Header" System. 

354. Typical "By-Pass" Piping System. 

355. General Arrangement of Steam and Exhaust Piping, Heyworth Building, 

Chicago, Illinois. 

356. General Arrangement of Piping, Manhattan Elevated Station, New York. 
357-8. Piping Arrangement at the Yonkers Power House of the New York Central. 

359. Overhead Piping of Boilers, Quincy Point Power Plant of the Old Colony Street 

Railway Company, Quincy Point, Mass. 

360. Main Stream Header and Branches, Grand Rapids, Grand Haven and Muskegon 

Railway Power House. 

361. Main Header and Branches, Des Moines City Railway Power House. 

362. Drop in Pressure in Steam Pipes of Various Diameters at Different Velocities. 

363. Diagrammatic Arrangement of Piping in the Webster Vacuum Heating System. 

364. Webster Vacuum Seal Valve. 

365. Automatic Vacuum Valve, Illinois Engineering Company. 

366. Diagrammatic Arrangement of Piping in the Paul Vacuum Heating System. 

367. Paul Exhauster. 

368. Paul Vacuum Valve. 

369. Powers Thermostat. 

370. Typical Diaphragm Valve. 

371. Diagram of Feed-Water Piping, Condensing Plant. 

372. Diagram of Feed- Water Piping, Non-Condensing Plant. 

373. Arrangement of Valves in Feed-Water Branches. 

374. Globe Valve, Screw-Top, Inside Screw. 

375. Globe Valve, Bolt-Top, Outside Screw. 

376. Gate Valve, Solid-Wedge, Screw-Top, Outside Screw. 

377. Gate Valve, Solid-Wedge, Bolt-Top, Inside Screw. 

378. Gate Valve, Split-Wedge, Bolt-Top, Inside Screw. 

379. Ludlow Angle Valve, Gate Pattern. 

380. Anderson Automatic Non-Return Valve. 

381. Crane Hydraulic Emergency Gate Valve. 

382. Anderson Triple-Duty Emergency Valve. 

383. Pilot Valve, Anderson Triple-Duty Emergency Valve. 

384. Types of Check Valves. 
385-7. Types of Blow-off Valves. 

388. Blow-off and Feed-Water Piping, South Side Elevated Railway Station, 

Chicago, Illinois. 

389. Dead-Weight Safety Valve. 

390. Common Lever Safety Valve. 

391. Consolidated Pop Safety Valve. 

392. Foster Back-Pressure Valve. 

393. Davis Back-Pressure Valve. 

394. Crane Atmospheric Relief Valve. 

395. Acton Atmospheric Relief Valve. 



ILLUSTRATIONS xxix 

Fig. 

396. Kieley Reducing Valve. 

397. Foster Pressure Regulator. 

398. Types of Foot Valves. 

Chapter XVI. — Lubricants and Lubrication. 

399. Oil Cup Lubrication. 

400. Nugent's Telescopic Oiler. 

401. Ring Oiler. 

402. Centrifugal Oiler. 

403. Pendulum Oiler. 

404. Simple Gravity Feed System. 

405. Low-Pressure Gravity Oil Feed. 

40G. Compressed-Air Oiling System at First National Bank Building, Chicago, 
Illinois. 

407. Leyland Automatic Oil Cup. 

408. Common Sight-Feed Hydrostatic Lubricator. 

409. Lunkenheimer Sight-Feed Lubricator. 

410. Central Hydrostatic Lubricating System. 

411. Rochester Forced-Feed Lubricator, Single Feed. 

412. Forced-Feed Cylinder Lubricator, Multi-feed. 

413. Siegrist System. 

414. Siegrist Sight-Feed Lubricator. 

415. White Star Oil Filter. 

416. Turner Oil Filter. 

Chapter XVII. — Finance and Economics — Cost of Power. 

417. Influence of Load Factor on Cost of Power. 

418. Cost of Power, Manufacturing Plant. 

418 a, b, c. Cost of Power in Large Central Stations. 

Chapter XVIII. — Testing and Measuring Instruments. 

419. Piston Water Meter. 

420. Disk Water Meter. 

421. Venturi Meter. 

422. St. John's Steam Meter. 
422a. Burnham Steam Meter. 

423. Different Forms of Manometer Draft Gauges. 

424. Bourdon Pressure Gauge. 

425. Bristol Recording Air Thermometer. 

426. Bristol Thermo-Electric Pyrometer. 

427. Element for Callendar Resistance Pyrometer. 

428. Wanner Optical Pyrometer. 

429. Fery Radiation Pyrometer. 

430. Orsat Apparatus. 

431. Arndt's Econometer. 

432. Ados C0 2 Recorder — Gas-Weighing Apparatus. 

433. Sarco C0 2 Recorder. 

434. Separating Calorimeter. 






xxx ILLUSTRATIONS 

Fig. 

435. Throttling Calorimeter. 
435a. Universal Calorimeter. 

436. Mahler Bomb Calorimeter. 

437. Parr Fuel Calorimeter. 

Chapter XX. — A Typical Steam Turbine Station. 

438. General Arrangement of Plant and Grounds. 

439. North Elevation of Building. 

440. General Plan of Boiler and Turbine Room. 

441. Section through Boiler and Turbine Room. 

442. Section through Boiler Room. 

443. General Plan Quarry Street Station. 

444. Side Elevation of Quarry Street Station. 

445. Sectional Elevation of Unit No. 4 Quarry Street Station. 

Chapter XXI. — A Typical Isolated Station. 

446. Plan of Ground Floor. 

447. Sectional Elevation through Line DD of Fig. 446. 

448. Cross Section through Line CC of Fig. 446. 

449. Longitudinal Section through BB of Fig. 446. 

450. Plan of Basement. 

451. Diagram of Switchboard Connections. 

Appendix B. — A. S. M. E. Rules for Conducting Boiler Tests. 

452. Ringlemann Smoke Chart. 

453. Graphical Log Boiler Test. 

Appendix C. — A. S. M. E. Rules for Conducting Steam Engine Tests. 

454-5. Rope Brakes. 

456. Alden Absorption Dynamometer. 

457. Indicator Card, Simple Engine. 

458. Indicator Card, Four- Valve Engine, Slow Speed. 

459. Indicator Card, Single- Valve Engine, High Speed. 
460-1. Temperature-Entropy Diagrams. 

Appendix H. — Mollier Diagram. 

462. General Outline of Mollier's Diagram. 

463. Complete Reproduction of Diagram. 






. 



STEAM POWER PLANT ENGINEERING 



CHAPTER I. 

ELEMENTARY STEAM POWER PLANTS. 

1. General. — An equipment for the generation of power is known 
as a station or plant. When equipped to generate electricity for the 
production of light or power it is known as an Electrical Station or 
Electric Light and Power Station. The term Heating Plant refers to 
a plant in which the heat energy of fuel is made available for heating 
purposes through the medium of steam or hot water. In general, 
plants or stations are designated according to the manner in which the 
energy of the fuel is utilized. 

When a station distributes power to a number of consumers more 
or less distant, it is called a Central Station. When the distances are 
very^ great, electrical current of high tension is frequently employed, 
and is transformed and distributed at convenient points through Sub- 
stations. A plant designed to furnish power or heat to a building or a 
group of buildings under one management is called an Isolated Station. 
For example, the power plant of an office building is usually called an 
isolated station. 

When the exhaust steam from the engines is discharged at approxi- 
mately atmospheric pressure, the plant is said to be operating non- 
condensing. When the exhaust steam is condensed, reducing the back 
pressure on the piston by the partial vacuum thus formed, the plant 
is said to operate condensing. 

When the exhaust steam may be used for manufacturing, heating, or 
other useful purposes, as is frequently the case in various manufac- 
turing establishments, and in large office buildings, it is usually more 
economical to run non-condensing, while power plants for electric 
lighting and power, pumping stations, air compressor plants, and others, 
in which the load is fairly constant and the exhaust steam is not 
required for heating, are generally operated condensing. 



2 STEAM POWER PLANT ENGINEERING 

2. Elementary Non-Condensing Plant. — Fig. 1 gives a diagrammatic 
outline of the essential elements of the simplest form of steam power 
plant. The equipment is complete in every respect and embodies 
all the accessories necessary for successful operation. Where a small 



r+l 




INJECTOR 



Fig. 1. Elementary Non-Condensing Plant. 

amount of power is desired at intermittent periods, as in hoisting 
systems, threshing outfits and traction machinery, the arrangement 
is substantially as illustrated. The output in these cases seldom 
exceeds 50 horse power and the time of operation is usually short, so 
the cheapest of appliances are installed, simplicity and low first cost 
being more important than economy of fuel. 

Such a plant has three essential elements: (1) The furnace, (2) the 
boiler and (3) the engine. Fuel is fed into the furnace, where it is 
burned. A portion of the heat liberated from the fuel by combustion 
is absorbed by the water in the boiler, converting it into steam under 
pressure. The steam being admitted to the cylinder of the engine 
does work upon the piston, and is then exhausted through a suitable 
pipe to the atmosphere. The process is a continuous one, fuel and 
water being fed into the furnace and the boiler in proportion to the 
power demanded. 



ELEMENTARY STEAM POWER PLANTS 3 

In such an elementary plant, certain accessories are necessary for 
successful operation. The grate for supporting the fuel during com- 
bustion consists of a cast iron grid or of a number of cast iron bars 
spaced in such a manner as to permit the passage of air through the 
fuel from below. The solid waste products fall through or are " sliced " 
through the grate bars into the ash pit, from which they may be removed 
through the ash door. The latter acts also as a means of regulating the 
supply of air below the grate. Fuel is fed into the furnace through 
the fire door, and when occasion demands, air may be supplied above the 
bed of fuel by means of this door. The combustion chamber is the space 
between the bed of fuel and the boiler heating surface, its office being to 
afford a space for the oxidation of the combustible gases from the solid 
fuel before they are cooled below ignition temperature by the com- 
paratively cool surfaces of the boiler. The chimney or stack discharges 
the products of combustion into the atmosphere, and serves to create 
the draft necessary to draw the air through the bed of fuel. Various 
forced draft appliances are sometimes used to assist or to entirely 
replace the chimney. The heating surface is that portion of the boiler 
area which comes into contact with the hot furnace gases, absorbs the 
heat and transmits it to the water. In the small plant, illustrated in 
Fig. 1, the major portion of the heating surface is composed of a number 
of fire tubes below the water line, through which the heated gases 
pass. The vol-ume above the water level is called the steam space. 
Water is forced into the boilers either by a feed pump or an injector. 
In small plants of the type considered, steam pumps are seldom em- 
ployed; the injector answers the purpose and is considerably cheaper. 
A safety valve connected to the steam space of the boiler automatically 
permits steam to escape to the atmosphere if an excessive pressure is 
reached. The water level is indicated by try cocks or by a gauge glass, 
the top of which is connected with the steam space and the bottom 
with the water space. Try cocks are small valves placed in the water 
column or boiler shell, one at normal water level, one above it, and one 
below. By opening the valves from time to time the water level is 
approximately ascertained. They are ordinarily used in case of acci- 
dent to the gauge glass. Fusible plugs are frequently inserted in the 
boiler shell at the lowest permissible water level. They are com- 
posed of an alloy having a low fusing point which melts when in con- 
tact with steam, thus giving warning by the blast of the escaping steam 
if the water level gets dangerously low. The blowoff cock is a valve 
fitted to the lowest part of the boiler to drain it of water or to discharge 
the sediment which deposits in the bottom. The steam outlet of a 
boiler is usually called the steam nozzle. 



4 STEAM POWER PLANT ENGINEERING 

The essential accessories of the simple steam engine include: A 
throttle valve for controlling the supply of steam to the engine; the 
governor, which regulates the speed of the engine by governing the steam 
supply; the lubricator, attached to the steam pipe, which is usually of 
the "sight feed" class and provides for lubrication of piston and 
valve. Lubrication of the various bearings is effected by oil cups 
suitably located. Drips are placed wherever a water pocket is apt to 
form in order that the condensation may be drained. The apparatus 
to be driven by the engine may be direct connected to the crank shaft 
or belted to the fly wheel or geared. 

In small plants of this type no attempt is made to utilize the exhaust 
steam except in instances where the stack is too short to create the 
necessary draft, in which case the, exhaust may be discharged up the 
stack. If the draft is produced by convection of the heated gases in 
the chimney, the fuel is said to be burned under natural draft; if the 
natural draft is assisted by the exhaust steam, the fuel is said to be 
burned under forced draft. The power realized from a given weight of 
fuel is very low and seldom exceeds 2\ per cent of the heat value of the 
fuel. The distribution of the various losses in a plant of, say, 40 horse 
power is approximately as follows: 

B.T.U. 

Heat value of 1 pound of coal 14,500 

Boiler and furnace losses, 50 per cent 7,250 

Heat of the steam, 50 per cent 7,250 

Heat equivalent of one horse power hour 2,545 

Heat used to develop one horse power hour (50 pounds steam per 
horse power hour, pressure 80 pounds gauge, feed water 62 degrees F.) 57,500 

Per cent. 

2 545 
Percentage of heat in the. steam, realized as work, ' .... 4.4 

O i ,OUU 

2 545 
Percentage of heat value of the coal realized as work, __ _ ' — rr-pz. 2.2 
• 57,500 -;- 0.50 



The power plant of the modern locomotive is very much like that 
illustrated in Fig. 1, the main difference lying in the type of boiler and 
engine. The entire exhaust from the engine is discharged up the 
stack through a suitable nozzle, since the extreme rate of combustion 
requires an intense draft. The engine is a highly efficient one compared 
to that in the illustration, and the performance of the boiler is more 
economical. In average locomotive practice about 6 per cent of the 
heat value of the fuel is converted into mechanical energy at the draw 
bar. In general, a non-condensing steam plant in which the heat of 
the exhaust is wasted is very uneconomical of fuel, eA^en under the 



ELEMENTARY STEAM POWER PLANTS 5 

most favorable conditions, and seldom transforms as much as 7 per cent 
of the heat value of the fuel into mechanical energy. 

3. Non-Condensing Plant. Exhaust Steam Heating. — Fig. 2 gives 
a diagrammatic arrangement of a simple non-condensing plant differ- 
ing from Fig. 1 in that the exhaust steam is used for heating pur- 
poses. This shows the essential elements and accessories, but omits 
a number of small valves, by-passes, drains, and the like for the sake 
of simplicity. The plant is assumed to be of sufficient size to warrant 
the installation of efficient appliances. Steam is led from the boiler 
to the engine by the steam main. The moisture is removed from the 
steam before it enters the cylinder by a steam separator. The moisture 
drained from the separator is either discharged to waste or returned to 
the boiler. The exhaust steam from the engine is discharged into the 
exhaust main where it mingles with the steam exhausted from the steam 
pumps. Since the exhaust from engines and pumps contains a large 
portion of the cylinder oil introduced into the live steam for lubricat- 
ing purposes, it passes through an oil separator before entering the 
heating system. After leaving the oil separator the exhaust steam 
is diverted into two paths, part of it entering the feed water heater where 
it condenses and gives up heat to the feed water, and the remainder 
flowing to the heating system. During warm weather the engine 
generally exhausts more steam than is necessary for heating purposes, 
in which case the surplus steam is automatically discharged to the 
exhaust head through the back pressure valve. The back pressure valve 
is, virtually, a large weighted check valve which remains closed when 
the pressure in the heating system is below a certain prescribed amount 
but which opens automatically when the pressure is greater than this 
amount. During cold weather it often happens that the engine exhaust 
is insufficient to supply the heating system, the radiators condensing 
the steam more rapidly than it can be supplied. In this case live steam 
from the boiler is automatically fed into the main heating supply pipe 
through the reducing valve. 

The condensed steam, and the entrained air which is always present, 
are automatically discharged from the radiators by a thermostatic valve 
into the returns header. The thermostatic valve is so constructed that 
when in contact with the comparatively cool water of condensation it 
remains open and when in contact with steam it closes. The vacuum 
pump or vapor pump exhausts the condensed steam and air from the 
returns header and discharges them to the returns tank. The small 
pipe S admits cold water to the vacuum pump and serves to condense 
the heated vapor, and at the same time supply the necessary make up 
water to the system. The returns tank is open to the atmosphere so 



STEAM POWER PLANT ENGINEERING 




ELEMENTARY STEAM POWER PLANTS 7 

that the air discharged from the vacuum pump may escape. From 
the returns tank the condensed steam gravitates to the feed water 
heater where its temperature is raised to practically that of the exhaust 
steam. The feed water gravitates to the feed pump and is forced into 
the boiler. There are several systems of exhaust steam heating in 
current practice which differ considerably in details, but, in a broad 
sense, are similar to the one just described. The more important of 
these will be described later on. 

During the summer months when the heating system is shut down, 
the plant operates as a simple non-condensing station and practically 
all of the exhaust steam, amounting to perhaps 60 per cent of the heat 
value of the fuel, is wasted. The total coal consumption, therefore, is 
charged against the power developed. During the winter months, 
however, all, or nearly all of the exhaust steam may be used for heating 
purposes and the power becomes a relatively small percentage of the 
total fuel energy utilized. The percentage of heat value of the fuel 
chargeable to power depends upon the size of the plant, the number 
and character of engines and boilers, and the conditions of operation. 
It ranges anywhere from 50 per cent to 100 per cent for the summer 
months and may run as low as 6 per cent for the winter months. This is 
on the assumption, of course, that the engine is debited only with the 
difference between the coal necessary to produce the heat entering the 
cylinder and that utilized in the heating system. 

4. Elementary Condensing Plant. — Under the most favorable con- 
ditions a non-condensing plant can never be expected to realize more 
than 7 per cent of the heat value of the fuel as power. In large non- 
condensing power stations the demand for exhaust steam is usually 
limited to the heating of the feed water, and as only 12 per cent or 
15 per cent can be utilized in this manner, the greater portion of the 
heat in the exhaust is lost. Non-condensing engines require from 20 to 
60 pounds of steam per hour for each horse power developed. On the 
other hand in condensing engines the steam consumption may be reduced 
to as low as 10 pounds per horse power hour. The saving of fuel 
is at once apparent. 

Fig. 3 gives a diagrammatic arrangement of a simple condensing 
plant in which the back pressure on the engine is reduced by condens- 
ing the exhaust steam. A different type of boiler from that in Fig. 1 
or Fig. 2 has been selected for the purpose of bringing out a few of 
the characteristic elements. The products of combustion instead of 
passing directly through fire tubes to the stack as in Fig. 1 are deflected 
back and forth across a number of water tubes, by the bridge wall and a 
series of baffles. After imparting the greater part of their heat to the 



STEAM POWER PLANT ENGINEERING 



fS* ..... t HH 



^^ 



■ numm'm^ 




ssssssss ^^^ ^^^ 



ELEMENTARY STEAM POWER PLANTS 9 

heating surface the products of combustion escape to the chimney 
through the breeching or flue. The rate of flow is regulated by a damper 
placed in the breeching as indicated. 

The steam generated in the boiler is led to the engine through the 
main header. The steam is exhausted into a condenser in which its 
latent heat is absorbed by injection or cooling water. The process 
condenses the steam and creates a partial vacuum. The condensed 
steam, injection water, and the air which is invariably present are 
withdrawn by an air pump and discharged to the hot well. In case the 
vacuum should fail as by stoppage of the air pump the exhaust steam 
is automatically discharged to the exhaust head by the atmospheric 
relief valve, and the engine will operate non-condensing. The atmos- 
pheric relief valve is a large check valve which is held closed by atmos- 
pheric pressure as long as there is a vacuum in the condenser. When 
the vacuum fails the pressure of the exhaust becomes greater' than 
that of the atmosphere and the valve opens. 

The feed water may be taken from the hot well or from any other 
source of supply and forced into the heater. In this particular case it is 
taken from a cold supply and upon entering the heater is heated by the 
exhaust steam from the air and feed pumps. From the heater it 
gravitates to the feed pump and is forced into the boiler. Various 
other combinations of heaters, pumps, and condensers are necessary 
in many cases, depending upon the conditions of operation. Feed 
pumps, air pumps, and in fact all small engines used in connection with 
a steam power plant are usually called auxiliaries. 

A well-designed station similar to the one illustrated in Fig. 3 is 
capable of converting about 10 per cent of the heat value of the fuel 
into mechanical energy. The various heat losses are approximately 
as follows: 

BOILER LOSSES. Per Cent. 

Loss due to fuel falling through the grate 2 

Loss due to incomplete combustion 2 

Loss due to heat carried away in chimney gases 23 

Radiation and other losses "8 

Total V. ~35 

Heat used by engines and auxiliaries (16 pounds of steam per ' ' ' 

I.H.P. hour, pressure 150 pounds, feed water 210° F.) . . . . 16,250 

Engine and generator friction, 5 per cent 812 

Leakage, radiation, etc., 2 per cent 325 

Total 17,387 

Heat equivalent of one electrical horse power 2,545 

Percentage of the heat value of the steam converted into electrical Cent. 

energy 14 .7 

Percentage of heat value of fuel converted into electrical energy 

2545 X 0.65 

17,387 ' •■ " y ' 5 



10 STEAM POWER PLANT ENGINEERING 

5. Condensing Plant with Full Complement of Heat-saving Appli- 
ances. — When fuel is costly it frequently becomes necessary for the 
sake of economy to reduce the heat wastes as much as possible. The 
chimney gases, which in average practice are discharged at a tem- 
perature between 450 degrees F. and 550 degrees F., represent a loss of 
20 per cent to 30 per cent of the total value of the fuel. If part of the 
heat could be reclaimed without impairing the draft the gain would 
be directly proportional to the reduction in temperature of the gases. 
Again, in some types of condensers all of the steam exhausted by the 
engine is condensed by the circulating water and discharged to waste. 
If provision could be made for utilizing part of the exhaust steam for 
feed-water heating, the efficiency of the plant could be correspondingly 
increased. In many cases the cost of installing such heat-saving devices 
would more than offset the gain effected, but occasions arise where they 
give marked economy. 

Fig. 4 gives a diagrammatic arrangement of a condensing plant in 
which a number of heat-reclaiming devices are installed. The plant is 
assumed to consist of a number of engines, boilers, and auxiliaries. 
Coal is automatically transferred from the cars to coal hoppers placed 
above the boiler, by a system of buckets and conveyors. These hoppers 
store the coal in sufficient quantities to keep the boiler in continu- 
ous operation for some time. From the hoppers the coal is fed 
intermittently to the stoker by means of a down spout. The stoker 
feeds the furnace in proportion to the power demanded and auto- 
matically rejects the ash and refuse to the ash pit. The ashes are 
removed from the ash pit when occasion demands, and are transferred 
to the ash hopper by the same system of buckets and conveyor which 
handles the coal. The ash hopper is usually placed alongside the coal 
hoppers and is not unlike them in general appearance and construction. 

The products of combustion are discharged to the stack through the 
flue or breeching. Within the flue is placed a feed-water heater called 
an economizer, the function of which is to absorb part of the heat from 
the gases on their way to the chimney. The heat reclaimed by the 
economizer varies widely with the conditions of operation and ranges 
between 5 per cent and 20 per cent. Since the economizer acts as a 
resistance to the passage of the products of combustion it is sometimes 
necessary to increase the draft either by increasing the height of the 
chimney or, as is the usual practice, by using a forced draft system. 

Part of the heat of the exhaust steam is reclaimed by a vacuum heater 
which is placed in the exhaust line between engine and condenser. 
For example, if the feed water has a normal temperature of 60 degrees F. 
and the vacuum in the condenser is 26 inches, the vacuum heater will 



ELEMENTARY STEAM POWER PLANTS 



11 




12 STEAM POWER PLANT ENGINEERING 

raise the temperature of the feed to say 120 degrees F. ; thereby 
effecting a gain in heat of approximately 6 per cent. If the feed supply 
is taken from the hot well the vacuum heater is without purpose, as the 
temperature of the hot well will not be far from 120 degrees F. 

Referring to the diagram, the path of the steam is as follows: From 
the boiler it flows through the boiler lead to the main header or equalizing 
pipe. From the main header it flows through the engine lead to the 
high-pressure cylinder. The exhaust steam discharges from the low- 
pressure cylinder through the vacuum heater and into the condenser. 
Part of the exhaust steam is condensed in the vacuum heater and gives 
up its latent heat to the feed water. The remainder is condensed by 
the injection water which is forced into the condenser chamber by the 
circulating pump. The condensed steam and circulating water gravitate 
through the tail pipe to the hot well. The air which enters the con- 
denser either as leakage or entrainment is withdrawn by the air pump. 
The steam exhausted by the feed pump, air pump, stoker engine, and 
other steam-driven auxiliaries is usually discharged into the atmospheric 
heater, which still further heats the feed water. 

Referring to the feed water, the circuit is as follows: The pump 
draws in cold water at a temperature of say 60 degrees F., and forces 
it in turn through the vacuum heater, the atmospheric heater, and the 
economizer into the boiler. The vacuum heater raises the temperature 
to 120 degrees F., the atmospheric heater increases it to 212 degrees F., 
and the economizer still further to about 300 degrees F. The heat 
reclaimed by this series of heaters is evidently the equivalent of that 
necessary to raise the feed water from 60 degrees F. to 300 degrees F., 
or approximately 24 per cent of the total heat supplied. In some 
plants the economizer only is installed, in others the economizer and 
atmospheric heater are deemed desirable, still others utilize all three. 
The distribution of the heat losses in a plant of this type operating under 
favorable conditions is approximately as follows: 



Per Cent. B.T.U. 
Delivered to engine, 15 pounds steam per I.H.P. hour; 

pressure 150 pounds, feed 60° F 100 17,482 

Delivered to feed pump 1.5 262 

Delivered to circulating pump • 1.5 262 

Delivered to air pump 2 349 

Delivered to small auxiliaries 1-5 262 

Loss in leakage and drips • • 0-5 87 

Engine and generator friction 5 



874 



Radiation and minor losses 1 175 

Total 19.753 



ELEMENTARY STEAM POWER PLANTS 13 

PerCent. B.T.U. 

Returned by vacuum heater 5.5 1,086 

Returned by atmospheric heater 7.9 1,560 

Returned by economizer 9.7 1,916 

Total 23.1 4,562 

Net heat delivered to engine in the form of steam to pro- 
duce one electrical horse power, 19,753—4,562 .... 15,191 

2545 
Percentage converted to electrical power - ■ .... 16.7 

Boiler efficiency 70 

Percentage of heat value of fuel necessary to produce one 

electrical horse power at switchboard rriaT — * * H • 7 

15,191 

The preceding figures give the results of very good practice. So 
much depends upon the size and character of the prime movers, the 
nature of the fuel, and the conditions of operation that no definite 
figure can be given for the percentage of heat converted to power in a 
given type of station. Six per cent represents good average practice 
in a non-condensing plant and 10 per cent in a condensing plant. 
Pumping stations operating continuously under full load have realized 
as much as 15 per cent of the total heat value of the fuel, but such 
performances are practically unobtainable in connection with steam- 
driven electrical power plants. Steam power plants as a class are very 
wasteful of fuel at the best. 

One of the best recorded performances to date (March, 1909) of a 
steam-electric power plant is that of the Pacific Light and Power 
Company at Redondo, Cal. When operating under regular commer- 
cial conditions approximately 14 per cent of the available heat of the 
fuel (crude oil) is realized as power at the switchboard. For a detailed 
description of the plant and the results of the acceptance tests, see 
Jour, of Elec. Gas and Power, Aug. 22, 1908. 



CHAPTER II. 

FUELS AND COMBUSTION. 

6. General. — The subject of fuels and combustion has been so 
extensively treated by various authorities that a comprehensive dis- 
cussion would be without purpose here, but in order to bring out more 
clearly the matter pertaining to the commercial design and operation 
of steam power plants a few of the essential elements will be briefly 
treated. 

The fuels used for steam making are coal, coke, wood, peat, mineral 
oil, natural and artificial gases, refuse products such as straw, manure, 
sawdust, tan bark, bagasse, and occasionally corn and molasses. 

In most cases that fuel is selected which develops the required power 
at the lowest cost, taking into consideration all of the circumstances 
that may affect its use. Occasionally the disposition of waste products 
is a factor in the choice, but such instances are uncommon. The boilers 
and furnaces are designed to suit the fuel selected. 

7. Classification of Fuels. — Fuels may be divided into three classes 
as follows: 

1. Solid fuels. 

a. Natural fuels: straw, wood, peat, coal. 

b. Prepared: charcoal, coke, peat and other briquettes. 

2. Liquid fuels. 

a. Natural: crude oils. 

b. Prepared: distilled oils, alcohol, molasses. 

3. Gaseous fuels. 

a. Natural: natural gas. 

b. Prepared: coal gas, water gas, producer gas, oil gas. 

8. Solid Fuels. — Solid fuels are of vegetable origin and exist in a 
variety of forms between that of a comparatively recent cellulose growth 
and that of nearly pure carbon as anthracite coal. They owe their forms 
to the conditions under which they were created or to the geological 
changes which they have undergone. With each succeeding stage the 
percentage of carbon increases. The chemical changes are approxi- 
mately as follows: 

14 



FUELS AND COMBUSTION 



15 



Substance. 



Pure cellulose . . . 

Wood 

Peat 

Lignite 

Brown coal .... 
Bituminous coal . . , 
Semi-bituminous coal. 

Anthracite 

Graphite 



Carbon. 


Hydrogen. 


Per Cent 


Per Cent 


44.44 


6.17 


52.65 


5.25 


59.57 


5.96 


66.04 


5.27 


73.18 


5.58 


75.06 


5.84 


89.29 


5.05 


91.58 


3.96 


100.00 





Oxygen. 



Per Cent 
49.39 
42.10 
34.47 
28.69 
21.14 
19.10 
6.66 
4.46 



All natural solid fuels contain more or less earthy or inorganic matter 
which is not combustible and therefore remains as ash, while the 
organic matter is consumed. Sometimes the percentage of ash is so 
great as to render them valueless for steam-making purposes. 

Origin and Formation of Fuel: Engng, Aug. 23, 1901; Am. Geol., Feb., 1899; Col. 
Guard, Sept. 10, 1897, Oct. 1, 1897, Jan. 14, 1898, Jan. 28, 1898, March 18, 1898, Sept. 
14, 1900; Ec. Geol., Oct., 1905; Eng. U.S., April 1, 1903; Ir. and Coal Td. Review, 
Feb. 4, 1898, July 13, 1906. 

9. Coal. — Coals are most satisfactorily classified according to the 
constituents of the combustible, as 





Fixed Carbon. 


Volatile Matter. 


Anthracite 

Semi-anthracite 

Semi-bituminous 

Bituminous, Eastern 

Bituminous, Western 


Per Cent 
97 to 92.5 
92.5 to 87.5 
87.5 to 75 
75 to 60 
65 to 50 
Under 50 


Per Cent 

3 to 7.5 

7.5 to 12.5 

12.5 to 25 

25 to 40 

35 to 50 


Lignite 


Over 50 



Classification of Coals: Am. Inst, of Min. Engrs., May, 1906, Sept., 1905; Mines and 
Minerals, Dec, 1906; Min. Rept., Apr. 26, 1906; Col. Guard, July 6, 1900; Power, 
Oct., 1906. 

10. Anthracite. — This is the most perfect form of coal and consists 
almost entirely of carbon; it contains very little hydrocarbon and burns 
with little or no smoke, is slow to ignite, burns slowly, and breaks into 
small pieces when rapidly heated. It requires a very large grate of 
about twice the surface necessary for bituminous coal. Large sizes 
may be burned in almost any kind of a furnace and with moderate draft. 
For small sizes a thinner bed has to be carried unless a strong draft is 
used. There is difficulty in keeping it free from air holes. When 



16 



STEAM POWER PLANT ENGINEERING 



possible, the coal should be at least six inches deep on the grates. On 
account of the large percentage of ash in the smaller size, the fire 
requires frequent cleaning. Anthracite does not require " slicing " 
and should be disturbed only when cleaning is necessary. 

Small Size Anthracite : Eng. and Min. Jour., Dec. 22, 1904. Heat Value of Anthra- 
cite, Small Sizes, and the Best Way of burning it : Col. Guard, Nov. 26, 1897. Anthra- 
cite Coal Mines and Coal Mining : Rev. of Rev., July, 1902. The Screening of Anthra- 
cite : Col. Guard, Sept. 20, 1901. Preparation of Anthracite : Mines and Min., March, 
1905. Anthracite Washeries : Am. Inst, of Min. Engrs., Nov., 1905. Anthracite 
Coal Fields of Pennsylvania: Min. Mag., March, 1905. Virginia Anthracite : Eng. 
News, Oct. 20, 1904; Mines and Min., March, 1906. Burning of Anthracite Culm of 
Poor Quality : Trans. A.S.M.E., 7-390. The Use of Electricity in Anthracite Mining : 
Eng. and Min. Jour., Feb. 2, 1907. 



Anthracite is classed and marketed according to sizes, the following 
division of mesh being adopted as standard at Wilkesbarre in 1891 : 



Egg 

Stove 

Chestnut 

Pea 

Buckwheat 

Rice 



coal must pass through 2| inch mesh and not through 2 inch 



2 
H 



Sizes over " pea coal " are prohibitive in price for steam power plant 
use and consequently the demand is limited to the smaller sizes. 

11. Semi-Anthracite. — This coal kindles more readily and burns 
more rapidly than anthracite. It requires little attention, burns freely 
with a short flame, and yields great heat with little clinker and ash. It 
is apt to split up on burning and waste somewhat in falling through the 
grates. It swells considerably but does not cake. It has less density, 
hardness, and metallic luster than anthracite, and can generally be dis- 
tinguished by its tendency to soil the hands while pure anthracite will 
not. 

12. Semi-Bituminous. — This coal is softer than semi-anthracite. 
It ignites easily and burns freely under a moderate draft. It gives an 
intense fire and is an excellent steam coal, but is apt to smoke con- 
siderably unless special provision is made to prevent it. 

13. Bituminous. — This coal contains a large and varying amount of 
volatile matter and requires careful firing to prevent smoke and clinker. 
Its physical properties vary widely, so much so that it is usually divided 
into three grades: 

1. Dry Bituminous coal is sometimes known as the free-burning 
bituminous. It is hard and dense, black in color, but brittle and 



FUELS AND COMBUSTION 17 

splintery. It ignites somewhat slowly, burns freely with a short, clean, 
bluish flame, little smoke, and without caking. 

2. Bituminous Caking coal swells up, becomes pasty, and fuses together 
in burning. It contains less fixed carbon and more volatile matter than 
the free-burning grades. Caking coal is rich in hydrocarbon and is 
particularly adapted to gas making. The flame is of a yellowish 
color. 

3. Long Flaming Bituminous coal is similar in many respects to the 
caking coal but contains a larger percentage of volatile matter. It is 
free burning, with a long, yellowish flame. It may be either caking or 
non-caking. 

COAL FIELDS OF THE UNITED STATES. 

Alabama: Mines and Minerals, May, 1901. 

Arkansas: Eng. and Min. Jour., Sept. 12, 1903, Oct. 28, 1905. 

Colorado: Min. Rept., Jan. 19, 1905, March 2, 1905; Jour. W. S. Engrs., Dec, 
1903 ; Mines and Minerals, May, 1905. 

Illinois: Min. Mag., March, 1905; Eng. and Min. Jour., Jan. 13, 1906. 

Indiana: Eng. Rec, Jan. 27, 1906; Min. Mag., March, 1905; Power, July and 
Aug., 1902. 

Indian Territory : Min. Rept., May 17, 1906. 

Kansas: Eng. and Min. Jour., Dec. 7, 1901. 

Kentucky : Col. Guard, Sept. 7, 1900. 

Michigan: Eng. and Min. Jour., June 30, 1900; Min. World, Feb. 9, 1907. 

Missouri: Am. Inst, of Min. Engrs., Jan., 1905. 

Montana: Min. Mag., March, 1905; Min. World, Nov. 24, 1906. 

Ohio : Min. Mag., March, 1905. 

Pennsylvania : Eng. and Min. Jour., Aug. 24, 1901; Trans. A.S.M.E., 4-217; Min. 
Mag., March, 1905; Pro. Eng. S. W. Penn., Jan., 1907. 

Texas : Mines and Min., Oct., 1905. 

Virginia: Mines and Min., March, 1906; Eng. News, Oct. 20, 1904. 

West Virginia: Eng. and Min. Jour., May 12, 1904. 

Wyoming : Min. World, May 6, 1905. 



GENERAL. 

Coal Mines of the United States : Peabody Atlas, A. Bement, Chicago, 111., Min. 
World, May 6, 1905. 

Coal Resources of the Pacific : Eng. Mag., May, 1902. 

Rocky Mountain Coal Fields: Min. Rept., Jan. 5, 1905; Jour. Asso. Eng. Soc, 
Dec, 1902. 

Coal Fields, U.S. Northwest : Rev. of Rev., Feb., 1903. 

Coal Fields, U.S. Southwest : Eng. and Min. Jour., Oct. 17, 1903. 

U.S. Coal Fields : Steam Boiler Practice, Wm. Kent. 

Report of Coal Testing Plant : U.S. Geological Survey, Washington, D.C. 



18 STEAM POWER PLANT ENGINEERING 

14. Lignite, or brown coal, is a substance of more recent geological 
formation than coal and represents a stage in development intermediate 
between coal and peat. Its specific gravity is low, 1.2, and when freshly 
mined it contains as high as 50 per cent of moisture. It is non-caking 
and gives a bright but slightly smoky flame. It is a low-grade fuel and 
is used where good coal is difficult to get. Exposure to the air causes 
it to split into fine pieces like air-slacked lime. It is very fragile and 
will not bear much handling in transportation. 

Eng. and Min. Jour., Nov. 22, 1902, Feb. 7, 1903; Mines and Min., July, 1901. 
Lignite of Northeastern Wyoming : Mines and Min., Feb., 1907. 

15. Peat, or Turf, is formed by the slow carbonization under water 
of a variety of accumulated vegetable materials. It is unsuitable for 
fuel until dried. Peat as ordinarily cut and dried is too bulky for com- 
mercial competition with coal, and is used only where coal is prohibitive 
in price. When properly prepared and compressed into briquettes peat 
is an excellent fuel. In Russia, Germany, and Holland peat briquettes 
have passed the experimental stage and several millions of pounds 
are manufactured annually. Peat is used but little in this country 
at present, but its possibilities are beginning to attract the attention 
of engineers. The proportion in which the various primary constit- 
uents exist in dried peat is approximately as follows: 

m 

Per Cent. 

Fixed carbon 35 

Volatile matter 60 

Ash 5 

Peat: Power, Sept., 1907; Engr. U.S., April 1, 1905; Min. World, Sept. 30, 1905; 
Col. Guard, Nov. 30, 1900 ; Mines and Min., July, 1901 ; Eng. and Min. Jour., Nov. 22, 
1902; Feb. 7, 1903, Eng. Rec, 52-191; Sci. Am. Sup., March 2, 1907; Elec. Engr., 
Lond., Dec. 6, 1907. 

Fuel Briquetting : Jour. Assn. Eng. Soc, Jan., 1906; Iron Age, April 19, 1906; 
Power, Dec, 1902; Eng. and Min. Jour., Nov. 8, 1902; Mines and Min., Oct., 1904; 
Power, March, 1905; Engr. U.S., May, 1905. 

16. Wood, Straw, Sawdust, Bagasse, Tanbark. — In certain locali- 
ties cord wood is still used as a fuel, but the steadily increasing values 
of even the poorest qualities are rapidly prohibiting its use for steam- 
generating purposes. Sawdust, shavings, tanbark and other waste 
products of wood are burned under boilers in situations where such 
disposition nets the best financial returns. Recent progress, however, 
in industrial chemistry shows that ethyl and wood alcohols and other 
valuable by-products can be cheaply made from sawdust, shavings, 



FUELS AND COMBUSTION 



19 



slashings and similar waste material, and it is not unlikely that their 
use for steaming purposes will be unheard of in a comparatively few 
years. Table 00 gives the physical and chemical characteristics of a 
number of woods. 

Wood as Fuel: Prac. Engr. U. S., Jan., 1910, p. 805; Power & Engr., June 30, 1908, 
p. 1015; Power, Dec, 1908, p. 772. 

Burning Sawdust: Prac. Engr. U. S., Jan.. 1910, p. 48; Power & Engr., April 7, 
1908, p. 536; Oct. 13, 1908, p. 613; Jour, of Elec, Oct., 1905. 



TABLE 00. 

PHYSICAL AND CHEMICAL PROPERTIES OF WOODS, STRAW AND TANBARK. 

(Prac. Engr. U. S., Jan., 1910.) 





o 

3 . 

3 -r. 

° £ 

^ 3 
<o p 

ll 


o 
O 

'53 


Equivalent Weight 
of Coal. 13,500 
B.T.U. 


6 

I* 

O 


o <v 
>> 

a 


,1 

0) CD 
X 

O 


= 3 

fP t-, 
I* 


< 


Calorific value, 
B.T.U. per 
Pound. 


>> 

o 

"3 
< 


Ash 


46 
43 
45 
42 
41 


3520 
3250 
2880 
3140 
2350 
2350 
1220 
4500 
3310 
3850 
3850 
3310 
1920 
2130 
2130 
1920 
3310 
1920 


1420 

1300 

1190 

1260 

940 

940 

580 

1800 

1340 

1560 

1540 

1340 

970 

1050 

1050 

970 

1340 

970 












5450 
5400 
5580 
5420 
5400 
5400 
6410 
5400 

.5460 
5460 
5400 
5460 
6830 

.6660 
6660 
6830 
5460 
6830 




Beech 

Birch 

Cherry 


49.36 
50.20 


6.01 
6.20 


42.69 
41.62 


0.91 
1.15 


1.06 
0.81 


Sharpless 
Hutton 












Sharpless 


Elm 35 












Hemlock 25 


























Sharpless 
Hutton 


Maple /Hard 1 49 
Oak Live 59 






















tt 


" 'White. 52 
" Red 45 


49.64 


5.92 


41.16 


1.29 


1.97 


Rankine 
Hutton 


Pine, White 25 
" Yellow 3fi 












a 












a 


Poplar 

Spruce 

Walnut 


36 
25 
35 
25 


49.37 


6.21 


41.60 


0.96 


1.86 


it 
it 












tt 


Willow .... 


49.96 


5.96 
6.06 


39.56 
41.30 


0.96 
1.05 


3.37 
1.80 


Rankine 


Average. '. 


49.70 














Straw. 
Wheat . . . 
Barley . . . 


o 

-1-3 

CO 




Water 
16.00 
15.50 
15.75 


35.86 
36.27 
36.06 


5.01 
5.07 
5.04 


37.68 
38.26 


0.45 
0.40 
0.42 


5.00 
4.50 
4.75 


5155 


Clark 

tt 


Average 


37.97 




Tanbark 

Dry 








51.80 


6.04 


40.74 




1.42 


9500 


Mvers 











Compressed. 



20 



STEAM POWER PLANT ENGINEERING 



Bagasse, or megass, is refuse sugar cane and is used as a fuel on the 
sugar plantations. Its heat value depends upon the proportions of 
fiber, molasses, sugar and water left after the extraction. The heat 
furnished by the different constituents is about as follows: Fiber, 
8325 B.T.U. per pound; sugar, 7223 B.T.U. per pound; and molasses, 
6956 B.T.U. per pound. Table gives the heat value of bagasse and 



TABLE 0. 

HEAT VALUES OF BAGASSE AND VARIATION WITH DEGREE OF EXTRACTION. 



Is 

15 


2 

.1 i 

o £ 

Ph 


Fiber. 


Sugar. 


Molasses. 


£6 
•Sh - 

a . 
_, "^ 

o o 
H 


Heat required to 

evaporate the Water 

present. B.T.U. 


cs P3 


Lbs. Bagasse required 
to equal lib. Coal of 
14,000 B.T.U. Cal- 
orific Power. 


Coal Equivalent per 

Ton of Cane. 

Pounds. 


6 
o 




Ph 




d 

^ CQ 
CM w 




d 

i gj 


of. 

Is 

t>H 


S3 

a En 
o> 

P. 


90 


0.00 

28.33 


100.00 
66.67 


8325 
5552 










8325 
5900 


339 


8325 
5561 


1.68 
2.52 


119 
119 


2465° 


85 


3.33 


240 


1.67 


116 


2236 


80 


42.50 


50.00 


4160 


5.00 


361 


2.50 


174 


4697 


509 


4188 


3.34 


120 


2023 


75 


51.00 


40.00 


3330 


6.00 


433 


3.00 


209 


3972 


611 


3361 


4.17 


120 


1862 


70 


56.67 


33.33 


2775 


6.67 


482 


3.33 


232 


3489 


679 


2810 


4.98 


120 


1732 


65 


60.71 


28.57 


2378 


7.15 


516 


3.57 


248 


3142 


727 


2415 


5.80 


121 


1612 


60 


63.75 


25.00 


2081 


7.50 


541 


3.75 


261 


2883 


764 


2119 


6.61 


121 


1513 


55 


66.12 


22.22 


1850 


7.78 


562 


3.88 


270 


2682 


792 


1890 


7.40 


121 


1427 


50 


68.00 


20.00 


1665 


8.00 


578 


4.00 


278 


2521 


815 


1706 


8.21 


122 


1350 


45 


69.55 


18.18 


1513 


8.18 


591 


4.09 


284 


2388 


833 


1555 


9.00 


122 


1284 


40 


70.83 


16.67 


1388 


8.33 


601 


4.17 


290 


2279 


849 


1430 


9.79 


123 


1222 


25 


73.67 


13.33 


1110 


8.67 


626 


4.33 


301 


2037 


883 


1154 


12.13 


124 


1077 


15 


75.00 


11.77 


980 


8.82 


637 


4.41 


307 


1924 


899 


1025 


13.66 


124 


1002 





76.50 


10.00 


832 


9.00 


650 


4.50 


313 


1795 


916 


879 


15.93 


126 


906 



variation with the degree of extraction, 
bagasse is shown in Fig. 4a. 



A typical furnace for burning 



Bagasse as Fuel: Prac. Engr. U. S., Jan., 1910; Engr. U. S., April 1, 1903; Jour. 
Assn. Engng. Soc, July, 1901; Engng., Feb. 18, 1910. 

Tanbark is usually quite moist; the amount of moisture varies with 
the leaching process used and averages around 65 per cent. In this 
condition it has a heat value of about 4300 B.T.U. per pound. If 
perfectly dry its heating power is approximately 6100 B.T.U. per pound. 
As in the case of all moist fuels, tanbark must be surrounded by heated 
surfaces of sufficient extent to insure drying out the fresh fuel as thrown 
on the fire. A successful furnace for burning tanbark is shown in 
Fig. 4b. 

Tanbark as a Boiler Fuel: Jour. A.S.M.E., Feb., 1910, p. 181; Jour. A.S.M.E., 
Oct., 1909, p. 951 ; Prac. Engr. U. S., Jan., 1910. 



FUELS AND COMBUSTION 



21 




22 



STEAM POWER PLANT ENGINEERING 




H 



FUELS AND COMBUSTION 



23 



17. Composition of Coal. — The uncombined carbon in coal is known 
as fixed carbon, while the hydrocarbons and other gaseous compounds 
which distill off on application of heat constitute the volatile matter. 
Refractory earths and moisture are found in varying quantities in 
different classes of coal and as they are incombustible tend to reduce 
the heat value of the fuel. That part of the fuel which is dry and free 
from ash is called the combustible, though the nitrogen and oxygen in the 
volatile matter are not actually combustible. The term " pure coal " 
has been suggested in this connection and is meeting with much favor. 
(Jour. W.S.E. 11-757.)* The various elementary constituents of a 
fuel must be determined by a careful chemical analysis, but in most 
cases it is only necessary to know the heating value, the per cent 
of moisture and ash and perhaps the per cent of sulphur. Table 1 
shows the composition of a number of American coals and gives a good 
idea of their chemical characteristics. 



TABLE 1. 

COMPOSITION OF TYPICAL AMERICAN COALS. 

(U.S. Geological Survey.) 





Anthracite. 


Semi-Bituminous. 




ii . 
. * 

tuo 3 

r 


3 a 

n * 

§ !^ 3 
X £ PS 


Is 

o 




III 


H 

- 3 .2 


Proximate analysis 

Water 

Volatile matter . . . 

Fixed carbon 

Ash 

Sulphur 

Ultimate analysis 

Carbon 

Hydrogen ....... 

Nitrogen 


1.97 
4.35 
86.49 
7.19 
0.64 

85.66 
2.78 
0.77 
2.87 

13963 


1.50 

7.84 

81.07 

9.59 

0.50 

83.20 
3 = 29 
0.95 
2.45 

i3954 


2.08 

7.27 

74.32 

16.33 

0.77 

75.21 
2.81 
0.80 
4.08 

12472 
12395 


0.65 

18.80 

75.92 

4.63 

0.57 

85.91 
4.58 
1.07 
3.24 

15190 
15104 


0.59 

18.52 

74.31 

6.58 

0.81 

81.05 
4.91 
2.15 
4.57 

14484 


1.28 
12.82 
73.69 
12.21 

2.01 

77.29 
3.74 
1.39 


Oxygen 

Calorific value 

Calorimeter 

Dulong's Formula 


3.36 

13406 
13831 



* H. J. Williams. 



* See also " Unit Coal and the Composition of Coal Ash, 
No. 37, Aug. 9, 1909. 



Univ. of 111. Bulletin 



24 



STEAM POWER PLANT ENGINEERING 





TABLE 1.- 


- Continued. 








Bituminous. 


Lignite. 




cu 

■s as . 

pq 


•ill 


ill 
gas 

GO 


o o -2 


CO << 


ft * § 

3 <U .3 

a 


Proximate analysis 
Water 


8.61 

32.40 

51.33 

7.66 

1.65 

68.14 
5.38 
1.34 

15.83 

12236 
12082 


3.15 
30.27 
56.17 
10.41 

1.26 

74.33 
4.96 
1.43 
7.61 

13406 
13371 


1.92 

36.56 

57.08 

4.44 

1.24 

78.31 
5.36 
1.85 
8.80 

14319 
14081 


5.31 
34.29 
36.24 
24.16 

4.30 

54.06 
4.57 
0.78 

12.13 

9848 
9929 


18.51 
35.33 
30.67 
15.49 
3.05 

47.34 
5.93 
0.66 

27.53 

8525 
8550 


10 86 


Volatile matter. . . 

Fixed carbon 

Ash 

Sulphur 


35.14 

46.90 

7.10 

64 


Ultimate analysis 

Carbon 

Hydrogen 

Nitrogen 

Oxygen 

Calorific value 

Calorimeter 

Dulong's Formula 


64.34 
5.73 
1.05 

21.14 

11435 
11299 



Report of Committee on Coal Analysis: Jour. Amer. Chem. Soc, 21-1116, 1899. 
Coal Testing : Mines and Min., Nov., 1905. Coal Testing : Col. Guard, April 12, 1906. 
Report of Government Coal-Testing Plant: Trans. A.S.M.E., Dec, 1905; Eng. and 
Min. Jour., Sept. 15, 1904; Mines and Min., June, 1906. Determination of Volatile 
Combustible Matter: Jour. Am. Chem. Soc, 21-1137, 1899; ibid., 28-1002, 1906. 
Proximate Constituents of Coal: Jour. Gas Light, Jan. 5, 1897. Ultimate Analy- 
sis: Col. Guard, Oct. 15, 1897. Analysis of Coal: Trans. A.S.M.E., 21-65. 
Sampling Coal: Chem. Engr., Nov., 1905; Am. Inst. Mech. Engrs., Sept., 1905. 
Illinois Coal Tests : Eng. News, Feb. 7, 1907. 

Sulphur in Coal : Eng. and Min. Jour., June 27, 1903 ; Mines and Min., Feb., 1906 ; 
Jour. Am. Chem. Soc, 24-852, 25-184, 26-1139. 

Ash: Relation between Composition and Fusibility in Coal Ash: Col. Guard, Oct., 
1897, March 18, 1898. Composition of Ash : Stromeyer, Steam Boilers, p. 11. 

18. Combustion. — By combustion is meant the chemical union of 
the combustible material of a fuel and the oxygen of the air. Theoretic- 
ally the process is a simple one, as it is only necessary to bring each 
particle of fuel previously heated to the kindling temperature in con- 
tact with the correct amount of oxygen and the combustion will be 
complete, the fuel oxidizing to the highest possible degree. In practice, 
however, the size and character of fuel, type of furnace, draft, impuri- 
ties in the fuel, and the mechanical difficulties affect combustion to 
such an extent as to render oxidation more or less incomplete. 

When heat is applied to coal combustion takes place in three 
separate and distinct stages: 

1. Absorption of heat. A fresh charge of fuel when thrown on a 
fire must first be brought to the kindling point in order that chemical 



FUELS AND COMBUSTION 25 

action may take place. The temperatures necessary to cause this union 
of oxygen and fuel are approximately as follows: 

Degrees F. Degrees F. 

Lignite dust 300 Cokes 800 

Sulphur 470 Anthracite lump 750 

Dried peat 435 Carbon monoxide 1211 

Anthracite dust 570 Hydrogen 1100 

Lump coal 600 

(Stromeyer, Marine Boiler Management and Construction, p. 93.) 

2. Vaporization of the hydrocarbon portion of the fuel and its 
combustion, the hydrocarbons consisting principally of olefiant gas, 
C 2 H 2 , marsh gas, CH 4 , tar, pitch, naphtha and the like. As these gases 
are driven off they become mixed with the entering air, and the carbon 
and hydrogen unite with the oxygen, forming carbon- dioxide, CQ 2 , and 
water vapor, H 2 0, respectively, and give up heat in doing so. If 
volatile sulphur is present it unites with the oxygen, forming sulphur 
dioxide, S0 2 , and also gives up heat, but its presence is objectionable, as 
the S0 2 , particularly, in the presenoe of moisture, attacks the metal 
of the furnace and boiler and causes rapid corrosion. If insufficient 
oxygen is present for complete oxidation, the carbon may burn to carbon 
monoxide, CO, and only a small portion of the available heat be liberated. 

3. Combustion of the solid or carbonaceous portion of the fuel. 
After the hydrocarbons have been driven off and oxidized the remain- 
ing solid matter is composed chiefly of carbon and ash. The carbon 
unites with the oxygen, forming carbon dioxide, carbon monoxide, or 
both, depending upon the completeness of combustion. The ash, of 
course, remains unconsumed. 

In commercial practicethe requirements for perfect combustion are a 
surplus of air, a thorough mixture of the fuel particles with the air, and 
a high temperature. The surplus of air above theoretical require- 
ments should be kept to a minimum, but even in the most scientifically 
designed furnace some excess is essential on account of the difficulty 
of properly mixing the gases, since the currents of combustible gases and 
air are apt to be more or less stratified. The products of combustion 
must be maintained at the kindling temperature until oxidation is 
complete, otherwise the carbon will be wasted as carbon monoxide or 
as smoke. The final products of combustion as exhausted by the 
chimney should consist only of carbon dioxide, water vapor, oxygen, 
nitrogen, and the oxides of impurities in the fuel. 

When the combustible elements unite with oxygen they do so in 
definite proportions, called the combining weights, which are always 
the same, and the union produces a fixed quantity of heat. Thus in the 



26 



STEAM POWER PLANT ENGINEERING 



combustion of carbon to CO, 12 pounds of carbon unite with 16 pounds 
of oxygen, forming 28 pounds of CO; hence one pound of carbon will 



form 



12 + 16 
12 



= 2J pounds of CO. 



The heat of combustion will be 4,451 B.T.U. per pound of carbon thus 

consumed. 

TABLE 2. 

DATA RELATIVE TO ELEMENTS MOST COMMONLY MET WITH IN CONNECTION 
WITH COMBUSTION OF FUEL. 



Substance. 



Acetylene. 

Air 

Ash 

Carbon 

Carbon .... 
Carbon dioxide 
Carbon monox- 
ide 

Hydrogen . . 
Marsh gas . . 
Nitrogen . . . 
Olefiant gas 
Oxygen 
Sulphur .... 
Water vapor 



Chemical 


Combining 


Formula. 


Weight. 


C 2 H 2 


26 


c " 


"u 


c 


12 


C0 2 


44 


CO 


28 


H 


1 


CH. 


16 


N 


14 


C 2 H 4 


28 





16 


S 


32 


H 2 


18 



Chemical Reaction. 



Oxygen. 



Air. 



Weight per Pound 

Substance in First 

Column. 



C 2 H 2 +5 0=2C0 2 +H 2 



c+o=co 

C + 2 = C0 2 



CO + O = C0 2 
2H + O = H 2 
CH 4 +4 0=C0 2 +2H 2 



C 2 H 4 +60=2C0 2 +2H 2 

s+2'6=so 2 " 



3.08 



1.33 
2.66 



0.57 

8 
4 

3.43 

1 



13.27 



5.78 
11.58 



2.47 
34.8 
17.4 

14.9' 

4.32 



Substance. 


Specific 
Heat at 
Constant 
Pressure. * 


Weight per 
Cubic Foot. 


Cubic Feet 
per Pound. 


Heat of Combustion 
B.T.U.t 


62° F. 


30" Hg.t 


Per Pound. 


Per Cubic Foot. 
62° F. 30" Hg. 


Acetylene 

Air 


0.2375 
0.2 

0.2169 

0.2426 

3.409 

0.593 

0.2438 

0.404 

0.2175 

0.1544§ 

0.48 


0.0685 
0.0759 

0.1159 

0.0737 

0.00527 

0.0421 

0.07376 

0.0737 

0.0843 

0.1686 


14.6 
13.16 

' 8. 627 

13.55 
189.8 
23.72 
13.55 
13.55 
11.86 
5.93 


21,465 

4,451 

14,544 

4,325 
62,032 
23,513 

21,344 




1470 


Ash 




Carbon 

Carbon 

Carbon dioxide 
Carbon monox- 
ide. 


320 


Hydrogen 

Marsh gas 

Nitrogen 

Olefiant gas . . . 

Oxygen 

Sulphur 

Water vapor. . . 


326 
991 

1570 



* Regnault. 



t Favre & Silberman. 



J Smithsonian Tables. 



so 2 . 



FUELS AND COMBUSTION 27 

Similarly in burning to C0 2 one pound of carbon will form 3f pounds 
of CO a and liberate 14,544 B.T.U. 

Table 2 gives the physical and chemical properties of the substances 
most commonly met with in connection with combustion. 

Combustion: Engr. U. S., Feb. 16, 1903, April 1, 1903, Jan. 1, 1907; Power, Feb. 
1905, Oct. 26, 1906; Eng. Mag., July, 1907; Prac. Engr. U. S., Jan., 1910. 

Simple Method of Determining Condition of Combustion : A. Bement, Jour. West. 
Soc. Eng., June, 1901. Effect of Altitude on Combustion : Power, Sept., 1906. Notes 
on Fuel Combustion in Power Plants : Elec. Engr., Lond., Aug. 10, 1906. 

19. Temperature due to Combustion. — If the heat liberated by the 
chemical union of any two elements is confined in such a way that no 
radiation can take place, the resulting increase in temperature of the 
products of combustion may be expressed 



t = K (i) 

ws 



in which 



t = the increase in temperature, degrees F. 
h = the heat liberated, B.T.U. per pound of combustible. 

s = the specific heat of the resulting gaseous products. 
w = weight of the gaseous products, pounds. 

Thus 1 pound of carbon in burning to C0 2 combines with 2f pounds 
of oxygen and forms 3f pounds of C0 2 . Taking the mean specific 
heat of C0 2 as 0.217, the resulting increase in temperature will be 

t = 14 ' 544 = 18,282 degrees F. 
3f X 0.217 ' & 

Such a temperature, of course, cannot be reached in practice, even 
with pure carbon and oxygen, on account of the radiation losses. 

In the ordinary furnace the oxygen is obtained from the atmosphere, 
which, neglecting a few impurities and minor elements, contains the 
following, mechanically mixed: 

Oxygen 23 parts by weight. 

Nitrogen 77 parts by weight. 

or 

Oxygen 21 parts by volume. 

Nitrogen 79 parts by volume. 

Hence in the combustion of one pound of pure carbon the products 

77 
of combustion contain not only 3| pounds of C0 2 , but — X 2f = 8.92 

pounds of nitrogen, giving a total of 3| -f 8.92 = 12.58 pounds. 



28 STEAM POWER PLANT ENGINEERING 

The nitrogen performs no useful office in combustion and passes through 
the furnace without change. It dilutes the products of combustion and 
reduces their temperature. Thus in the specific case taken above the 
theoretical increase in temperature will be 

14 544 
' " 12.58 X 0.236 " 49 °° de ^ S F " 

The mean specific heat, 0.236, of the products of combustion is determined as 
follows: 

Heat necessary to raise 8.92 lbs. of nitrogen 1 degree is 8.92 X 0.2438 (specific 
heat of nitrogen) = 2.1747 B.T.U. 

Heat necessary to raise 3.66 lbs. of C0 2 1 degree is 3.66 X 0.217 (specific heat of 
C0 2 )= 0.7942 B.T.U. 

Total heat = 2.1747 + 0.7942 = 2.9689 B.T.U. 

2 9689 
Mean specific heat = ' = 0.236. 

o.y^ + o.oo 

Evidently for maximum temperature this dilution should be kept 
as low as possible. Unfortunately, in practice, a perfect union of fuel 
and air in theoretical proportions is almost impossible, and to insure 
complete combustion an excess of air is necessary. This still further 
reduces the temperature of the products of combustion. For example, 
the complete oxidation of one pound of carbon requires 11.5 pounds of 
air. If 20 pounds of air are supplied per pound of carbon, an average 
figure in good practice, the theoretical increase in temperature will be 

' = 21^^36 " 2934 degreeS R 

In the preceding computation the specific heat of the constituent 
gases is assumed to be constant for all temperatures. Recent experi- 
ment seems to show that the specific heat increases at high tempera- 
tures, though actual values are not yet available. 

20. Air required for Combustion. — The quantity of air required 
for complete combustion may be approximately determined from the 
ultimate analysis. Thus 

A = 34.56^+H-^, (2) 

in which 

A = weight of air required per pound of fuel. 

C, H, and O = proportional part by weight of carbon, hydrogen, and 
oxygen in the fuel. 

When the flue-gas analysis is known, the air actually supplied may be 
approximately determined as in the following example: 

Example: Required the weight of air supplied per pound of coal with 
fuel and flue gas analysis as follows: 



FUELS AND COMBUSTION 



29 



Coal analysis (Illinois bituminous, mine run) : 

Carbon 68.14 Nitrogen 1.34 

Oxygen 15.83 Ash 7.66 

Hydrogen 5.38 Sulphur 1.65 

Flue gas analysis, cubic feet per 100 cubic feet of dry gas (tempera- 
ture, 62° F., barometer 30 inches): 
C0 2 14 
CO 0.5 
O 6.0 
S0 2 (not determined, since its weight is 
practically negligible). 
The weights may be determined from the densities given in Table 2. 





Volume. 


Density. 


Weight. 


C0 2 


CO 


14 
6 
0.5 


0.1159 
0.0843 
0.0737 


1.6226 
0.5058 
0.0368 



These weights can be subdivided into those of their constituents; 
thus the C0 2 contains T 3 T of carbon and T 8 T of oxygen, and the CO, f of 
carbon and .$ of oxygen. 

T 8 T X 1.6226 = 1.1800 T 3 T X 1.6226 = 0.4426 

f X 0.0368 = 0.0210 f X 0.0368 = 0.0158 

0.5058 Pounds of carbon 0.4584 
Pounds of oxygen 1.7068 

Weight of oxygen per pound of carbon = ' 
Since 23% of air by weight is oxygen, 



3.72. 



16.2. 



3 72 
Weight of air per pound of carbon = j^z 

Weight of air supplied per pound of coal is 

for the carbon, 0.6814 X 16.2 = 11.04 lbs. 

for the hydrogen, 34.56 (o.0538 - Qjj§i^ = 1.17 lbs. 
total V ' 12.21 lbs. 

The theoretical weight of air for combustion would have been 

0.1583\ = 



11.58 X 0.6814 + 34.56 0.0538 - 



.56 A 



8 / 



9.06 lbs. 



The percentage of air excess 



12.21-9.06 
100 9^6 = 35 - 



The weight of air actually supplied may be closely approximated 
by the following equation : 

A=3 - 032 (cofco) XC ' (2a) 



30 



STEAM POWER PLANT ENGINEERING 



in which A = the weight of air supplied per pound of fuel. 
N, CO, C0 2 = percentages by volume of nitrogen, 

carbon monoxide, and carbon dioxide in the flue gas. 
C = the proportional part by weight of carbon in the fuel. 
The per cent of excess air supplied per pound of fuel may be con- 
veniently determined from the relationship 

Air actually required _ N 

Air theoretically supplied ~~ N — 3.782 0* 
N and O are respectively, by volume, the proportional parts of the 
nitrogen and oxygen in the flue gas. The free oxygen is due to the air 
supplied and not used. This oxygen was accompanied by 3.782 times 
its volume of nitrogen. (N — 3.782) represents the nitrogen content 
in the air actually required for combustion. Hence, N-s-(N — 3.782) 
is the ratio of the air supplied to that required. 

Table 2a gives the values of the ratio corresponding to various per- 
centages of C0 2 + CO and C0 2 + CO + O. 

TABLE 2a. 

RATIO OF TOTAL AIR SUPPLIED TO THAT THEORETICALLY REQUIRED 

FOR VARIOUS ANALYSES OF FLUE-GASES. 

N 



Ratio 



N- 3.7820 





N=79. 


N = 79.5. 


N=80. 


N=80.5. 


N=81. 


N=81.5. 


N=82. 


co 2 +co. 


CO2+CO 


co 2 +co 


COa+CO 


co 2 +co 


COg+CO 


COa+CO 


COg + CO 




+ 0=21. 


+0=20.5. 


+ 0=20. 


+0=19.5. 


+ 0=19. 


+0=18.5. 


+ 0=18. 


21 


1.02 














20 


1.05 


"l.00 


1.00 










19 


1.11 

1.17 


1.08 
1.14 


1.05 
1.10 


i!o2 

1.08 


"l.OO 
1.05 






18 


1^02 


i!oo" 


17 


1.24 


1.20 


1.17 


1.13 


1.10 


1.07 


1.05 


16 


1.32 


1.27 


1.23 


1.20 


1.16 


1.13 


1.10 


15 


1.40 


1.35 


1.31 


1.27 


1.23 


1.19 


1.16 


14 


1.51 


1.45 


1.39 


1.35 


1.30 


1.26 


1.23 


13 


1.62 


1.55 


1.50 


1.44 


1.39 


1.34 


1.30 


12 


1.76 


1.68 


1.61 


1.54 


1.49 


1.43 


1.38 


11 


1.92 


1.82 


1.74 


1.66 


1.60 


1.53 


1.48 


10 


2.11 


2.00 


1.90 


1.81 


1.72 


1.65 


1.59 


9 


2.35 


2.21 


2.08 


1.97 


1.88 


1.79 


1.71 


8 


2.65 


2.47 


2.31 


2.18 


2.06 


1.95 


1.86 


7 


3.03 


2.80 


2.59 


2.44 


2.27 


2.14 


2.03 


6 


3.55 


3.22 


2.96 


2.74 


2.54 


2.38 


2.24 


5 


4.27 


3.81 


3.44 


3.14 


2.89 


2.68 


2.50 


4 


5.37 


4.65 


4.11 


3.68 


3.34 


3.05 


2.83 


3 


7.23 


5.97 


5.10 


4.45 


3.96 


3.56 


3.25 


2 


11.06 


8.34 


6.71 


5.63 


4.85 


4.27 


3.82 


1 


23.51 


13.83 


9.83 


7.64 


6.27 


6.12 


4.64 



The weight of air per pound of combustible as indicated by the per 
cent of C0 2 in the flue gas is given in Table 3. These figures are only 
approximate but are sufficiently accurate for many practical purposes. 



FUELS AND COMBUSTION 



31 



TABLE 3. 

WEIGHT OF AIR PER POUND OF COMBUSTIBLE AS INDICATED BY THE PER- 
CENTAGE OF C0 2 IN THE FLUE GAS. 





Pounds of 


Per Cent of 


Pounds of 


Per Cent of 


Pounds of 


Per Cent of C0 2 . 


Air. 


co 2 . 


Air. 


co 2 . 


Air. 


21 


12 


14 


18 


7 


36 


20 


12.6 


13 


19.4 


6 


42 


19 


13.3 


12 


21 


5 


50.5 


18 


14 


11 


22.9 


4 


63 


17 


14.8 


10 


25.2 


3 


84 


16 


15.7 


9 


28 


2 


126 


15 


16.8 


8 


31.5 


1 


210 



21. Calorific Value of Coal. — The heat liberated by the combustion 
of unit weight of fuel is called the calorific value of the fuel. The most 
rational way of determining the heat of combustion is to burn a 
weighed sample of coal in an atmosphere of oxygen in a suitable 
calorimeter. An alternative method is to calculate the heat of com- 
bustion from the chemical analysis. An analysis which determines 
the per cent of fixed carbon, volatile matter, moisture, and ash is called 
the proximate analysis, while one which reduces the fuel to its ele- 
mentary constituents of carbon, hydrogen, nitrogen, sulphur, moisture, 
and ash is called the ultimate analysis. The proximate analysis is 
comparatively easy to make and gives the general characteristics of a 
fuel. It is made by subjecting the sample to a moderate temperature 
to expel the moisture, then to a higher temperature until the volatile 
matter is driven off, and finally to a very high temperature which drives 
off all carbon as carbon dioxide and leaves the ash as a residue. By 
weighing the residue at the end of each operation the various percent- 
ages may be computed. For method of making proximate analysis, see 
" Report of Committee on Coal Testing," Journal of the American 
Chemical Society, Vol. 21, p. 1116. 

The heating value of certain classes of coals may be estimated from 
the proximate analysis. Thus for Illinois coals with ash content under 
10 per cent, R. W. Hunt & Co. deduced the formula 



h = 14,544 C + 16,515 V - 10,000 A, 



(3) 



in which 



h = B.T.U. per pound of coal. 

C, V, and A = the proportional content of fixed carbon, volatile 
matter, and ash. 



32 STEAM POWER PLANT ENGINEERING 

When ash lies between 10 and 15 per cent, the formula will be more 
accurate if written 

h = 14,544 C + 16,515 V + 354 A - 1635 (4) 

Kent, Goutal, and other authorities have deduced formulas based on 
the proximate analysis which agree closely with the calorimetric deter- 
minations for many coals, but a calorimetric determination is necessary 
whenever exact results are required. 

The formula most commonly used in calculating the heating value of 
a fuel is based on the ultimate analysis and is known as Dulong's 
formula. Thus 



h = 14,600 C + 62,000 (h - ^ + 4,000 S, 



(5) 



in which 



h = heating value in B.T.U. per pound of coal. 
C, H, O, and S refer to the proportions of carbon, hydrogen, oxygen, 
and sulphur, respectively. 

The calorimetric determination is more readily made and is usually 
more reliable than the calculation from the ultimate analysis. Table 1 
gives the proximate and ultimate analyses and the calorimetric and cal- 
culated heat values for a number of American coals. 

Calorific Value of Fuels: Poole, Engr. U.S., April 1, 1903; P. Mahler, Jour. 
Frank. Inst., Jan., 1905; A. Adams, Jour. Soc. Chem. Ind., Oct. 31 and Nov. 30, 
1901 ; Power, Oct. 1, 1906; Mines and Min., May, 1902; Trans. Am. Inst. Min. Engrs., 
27-259; Kent, Steam Boiler Economy, Chap. V; Report of Government Coal Testing 
Plant at St. Louis. 

Calorimetry : Berthier Method of Coal Calorimetry : Trans. A.S.M.E., 21-304. 
Comparison of Calorimeters: Jour. Soc. Chem. Ind., 22-1230, 1903; 23-704. Im- 
proved Form of Thompson Calorimeter : Jour. Soc. Chem. Ind., 25-409, 1906. Parr's 
Calorimeter: Chem. Engr., Feb., 1907. Coal Calorimeters : Trans. A.S.M.E., 14- 
816, 16-1040. Mahler Bomb: Calorific Power of Fuels, Poole; Trans. Am. Inst. 
Min. Engrs., 27-259, Engr. U.S., Jan. 1, 1907, p. 68. 

22. Heat Losses in Burning Coal. — The function of the boiler is the 
absorption of the heat liberated by the combustion of the fuel. In 
practice from 50 per cent to 85 per cent of this heat is utilized in mak- 
ing steam, depending upon the conditions of operation, and the remainder 
is wasted. Complete utilization of the heat of combustion is impossible. 
A boiler and furnace which utilizes 80 per cent of the heat in the fuel 
is exceptional and an average figure for very good practice is not 



FUELS AND COMBUSTION 33 

far from 75 per cent. The various losses may be summed up as 
follows: 

1. Loss in the dry chimney gases. 

2. Loss due to incomplete combustion. 

3. Loss of fuel through the grate. 

4. Superheating the hygroscopic moisture in the air. 

5. Moisture in the fuel. 

6. Loss due to the presence of hydrogen in the fuel. 

7. Unburned fuel carried beyond the combustion chamber in the 
form of soot or smoke. 

8. Radiation and minor losses. 

Fuel Losses: G. H. Barms, Cassier's Mag., Aug., 1907 ; Power, Sept. 14, 1909, p. 439. 

23. Loss in the Dry Chimney Gases. — The most serious loss is the 
heat carried away in the chimney gases. This may be expressed 

h = (W + 1) (T - t) S, (6) 

in which 

h = loss in B.T.U. per pound of combustible. 
W = weight of air supplied per pound of combustible. 
T = chimney temperature, degrees F. 

t = temperature of the air entering furnace. 

S = specific heat of the gases. (This may be taken as 0.24 for 
most purposes.) 

This loss is unavoidable and is seldom less than 15 per cent of the total 
heat supplied. In average good practice it is not far from 20 per cent. 
Even in the ideal case when pure carbon burns to C0 2 and the air 
supply is theoretically correct, the loss is appreciable. 

For example, suppose the temperature of the air is 50 degrees F., the 
temperature of the flue gas 450 degrees F., and the air supply 11.5 
pounds per pound of combustible, the heat carried away by the chim- 
ney will be 

12.5 X (450 - 50) X 0.24 = 1200 B.T.U., 

which is 1200 ■*■ 14,500 = 8.3 per cent of the total heat supplied. 

As a matter of fact, considerable excess of air is necessary for com- 
plete combustion, depending upon the size and nature of the fuel, 
thickness of the fire, variation in resistance through the fuel, intensity 
of draft, and character of grate. Practice shows a minimum excess of 
50 per cent in well-proportioned furnaces working under ideal conditions 
with an average not far from 100 per cent. 

Table 4 indicates the magnitude of these losses for different chimney 
temperatures and weights of air per pound of combustible. 



u 



STEAM POWER PLANT ENGINEERING 



TABLE 4. 

HEAT CARRIED AWAY BY THE DRY CHIMNEY GASES PER POUND OF 

COMBUSTIBLE. 



6 

3 

tn 

3 

s 
6 

O 

■a 
& 

o 

p. 

(-1 

< 

'o 

'O 

c 

§ 




Temperature of Chimney Gases. Deg. Fahr. 


300° 


350° 


400° 


450° 


500° 


550° 


600° 


650° 


12 

* 


750 

5.2 


905 

6.2 


1060 

7.3 


1216 

8.7 


1370 

9.5 


1528 

10.5 


1683 

11.6 


1840 

12.7 


15 


865 

6 


1112 

7.6 


1305 

9.1 


1498 

10.3 


1679 

11.6 


1880 
13.0 


2072 

14.3 


2262 

15.6 


18 


1004 

7.2 


1321 

9.1 


1550 

10.7 


1778 
12.2 


2010 

13.9 


2235 

15.4 


2460 

17 


2692 

17.9 


21 


1266 

8.7 


1530 

10.5 


1785 

12.3 


2060 

14.2 


2320 

16 


2582 
17.8 


2846 

19.5 


3118 

21 


24 

27 
30 
33 
36 
39 
42 


1440 

9.9 


1740 

12 


2040 

14 


2340 

16.1 


2640 

18.2 


2940 

20.3 


3240 

22.4 


3540 

24.4 


1611 

11.1 


1950 

13.5 


2281 

15.7 


2620 

18.1 


2958 

20.4 


3291 

22.7 


3628 

25 


3962 

27.4 


1785 

12.4 


2160 

14.9 


2530 

17.4 


2900 

20 


3270 

22.6 


3641 

25 


4016 

27.8 


4396 

30.4 


1957 

13.5 


2362 

16.3 


2779 

19.2 


3180 

22 


3589 

24.7 


4000 

27.6 


4405 

30.5 


4820 

33.2 


2130 

14.7 


2579 

17.8 


3020 

20.8 


3461 

23.9 


3910 

27 


4350 

30 


4798 

33 


5290 

36.6 


2300 

15.9 


2781 

19.2 


3261 

22.5 


3743 

25.8 


4220 

29.2 


4700 

32.4 


5180 

35.7 


5670 

39 


2479 

17.1 


2999 

20.6 


3508 

24.7 


4023 

27.7 


4540 

31.3 


5052 

34.8 


5570 

39.4 


6100 

42 



* Theoretical requirement. 

Large type gives the loss in B.T.U. per pound of combustible. 
Small type gives the per cent loss, assuming a calorific value of 14,500 B.T.U. 
per pound of combustible. 

From the table it will be seen that this loss is approximately 12 per 
cent of the total heat supplied when the excess of air is 50 per cent and 
16 per cent when the excess is 100 per cent. Flue-gas temperature 
assumed to be 450 degrees F. Flue temperatures less than 450 degrees F. 
are seldom experienced except in connection with economizers, and an 
average figure is about 500 degrees F. 



FUELS AND COMBUSTION 



35 



The weight of air supplied per pound of combustible is most con- 
veniently determined by the percentage of C0 2 in the flue gas. Thus 
for the complete oxidation of pure carbon the resulting flue gases 

TABLE 5. 

LOSS DUE TO INCOMPLETE COMBUSTION OF CARBON TO CARBON 

MONOXIDE. 



a5 

s 

"o 

> 

>> 
£> 

CO 

O 

9 

fl 

s 

43 

a 

o 
o 

c 
o 

Pi 




Per Cent of C0 2 in the Flue Gas by Volume. 


6 


8 


10 


12 


14 


16 


0.2 


328 

2.2 


248 
1.7 


199 

1.3 


168 

1.1 


144 

1 


126 

0.8 


0.4 


635 

4.3 


484 

3.3 


390 

2.6 


327 

2.2 


282 
1.9 


248 

1.7 


0.6 


925 

6.3 


709 

4.8 


575 

3.9 


474 

3.2 


417 

2.8 


367 

2.5 


0.8 
1.0 


1192 

8.1 


923 

6.3 


750 

5.1 


635 

4.3 


549 

3.7 


495 

3.4 


1494 

10.2 


1128 

7.7 


923 

6.3 


780 
5.3 


676 

4.6 


596 

4.1 


1.2 


1690 

11.5 


1321 

9 


1085 

7.4 


923 

6.3 


801 

5.4 


708 
4.8 


1.4 


1920 

13.1 


1512 

10.3 


1248 

8.5 


1061 

7.2 


924 

6.3 


819 

5.6 


1.6 


2104 

14.3 


1693 

11.5 


1400 

9.5 


1193 

8.1 


1040 

7.1 


924 

6.3 


1.8 


2340 

16 


1865 

12.7 


1549 

10.5 


1321 

9.0 


1151 

7.8 


1025 

7 


2.0 


2537 

17.2 


2030 

13.8 


1690 

11.5 


1450 

9.9 


1270 

8.6 


1129 

7.7 



Large type gives the loss in B.T.U. per pound of carbon. Small type gives the 
per cent loss, assuming a calorific value of 14,650 B.T.U. per pound of carbon. 

should consist of carbon dioxide and nitrogen only, and in the ratio 
by volume of 21 to 79; therefore 21 per cent of C0 2 in the flue gas 
is indicative of complete combustion and theoretical air supply. In 
other words, the ratio by volume of C0 2 to N after complete com- 
bustion is practically the same as the ratio of the oxygen to the nitrogen 
in the air before combustion. Table 3 gives the approximate weight 



36 



POWER PLANT ENGINEERING 



of air used per pound of combustible for different percentages of C0 2 
in the flue gas. 

In practice, 15 per cent is all that can be expected under the best 
conditions, with an average between 10 per cent and 12 per cent. Any- 
thing less than 10 per cent shows an excessive amount of air supplied. 
Traveling grates, unless carefully operated, are apt to show as low as 
5 per cent of C0 2 . 

24. Loss due to Incomplete Combustion. — If the volatile gases are 
not completely oxidized, as when the air supply is insufficient or the 
mixture of air and gases is not thorough, some of the carbon may 
escape as CO. The presence of even a small amount of CO in the flue 
gas is indicative of a very appreciable loss, as will be seen from Table 5. 
Carbon monoxide is a colorless gas, and its presence in the chimney 
gases cannot be detected by the fireman, consequently the absence of 
smoke is not an infallible guide for perfect combustion. This loss may 
be expressed 

, _ „ 10,150 CO 

K ~ ° X (C0 2 + CO)' (7) 

in which h 2 = the loss in B.T.U. per pound of carbon, C0 2 and CO 
are percentages by volume of the flue gases and C is the proportional 
part of carbon in the combustible. 



15 


















































































































14 
13 


















































































C 


J-i 






























12 
11 
10 
9 
S 
7 
6 
5 
4 
















































































Relation of Gas Composition in Rear 

Combustion Chamber To Temperature 

at Same Place 
































































































































































































































■2i 




















































































3 
2 

1 












































































































































_C 


u_ 























































































0.5 o 

0.4 t 

0.3 J 

0.2 O 

o.i a 



1900 2000 



2100 2200 2300 2400 2500 2600 

Combustion Chamber Temperature.Deg.Fah. 



2700 



Fig. 5. Relation of Gas Composition in Combustion Chamber to Temperature. 



This loss, however, may be wholly avoided in a properly designed 
and carefully operated furnace. In fact the loss from this cause is 
often exaggerated and seldom exceeds 2 per cent of the total heat value 
of the fuel except during the few moments following the replenishing 



FUELS AND COMBUSTION 37 

of a burned-down fire with fresh fuel or when the supply of air is checked 
to meet a sudden reduction in load. In improperly designed furnaces 
in which the volatile gases are brought into contact with the cooler 
boiler surface before combustion is complete, the carbon monoxide may 
be reduced in temperature below its ignition point and consequently 
will fail to combine with the oxygen. In such a case the loss may 
prove to be a serious one. Fig. 5 shows the relation between the com- 
position of the products of combustion in the rear combustion chamber 
of a 250-horse-power Heine boiler, hand fired, and the temperature at the 
same place. (For an extended discussion of this subject see Jour. 
West. Soc. Engrs., June, 1907, p. 285.) 

25. Loss of Fuel through Grate. — The refuse from a fuel is that 
portion which falls into the pit in the form of ashes, unburned or partially 
burned fuel, and cinders. The loss from this cause depends upon the 
size of the fuel, the width of opening in the grate bars, and the type of 
grate. Coal which necessitates frequent slicing is apt to give greater 
loss than a free-burning coal. Under good conditions of operation it 
ought not to exceed 2 per cent of the total heat value of the fuel. 
In traveling grates in which a large percentage of the fine fuel falls 
through the front end of the grate a special hopper is ordinarily installed 
in the ash pit which reclaims most of it. (See Fig. 99.) 

Loss of Fuel in Ashes : Power, March, 1905. Experiments on Fuel Value of Bitumi- 
nous Coal Ashes: Technology Quarterly, Dec, 1905. Coal Ash: Jour. Soc. Chem. 
Ind., Jan. 15, 1904. 

26. Superheating the Moisture in the Air. — The loss due to this 
cause is a minor one, though on hot, humid days it may be appreciable. 
This loss may be expressed 

h 3 = 0.48 M(T - t)* (8) 

in which h 3 = B.T.U. lost per pound of combustible. 

M = weight of moisture introduced with the air per 
pound of combustible. 
t = temperature of air entering the furnace, degrees F. 
T = temperature of chimney gases, degrees F. 

27. Moisture in the Fuel. — Moisture in the fuel represents an appre- 
ciable loss in economy if present in large quantities, since the heat 
necessary to evaporate it into superheated steam at chimney temperature 
is lost. Firemen occasionally wet the coal to assist coking or to reduce 

* The latest accepted value for the mean specific heat of water vapor at atmos- 
pheric pressure is 0.46 in place of 0.48. 



38 STEAM POWER PLANT ENGINEERING 

the dust, but moisture thus added necessarily reduces the furnace 
efficiency. The loss due to this cause is expressed: 

h 4 = M [(212 - t) + 966* + 0.48f (T - 212)], (9) 

in which h 4 = B.T.U. lost per pound of combustible. 

M = weight of moisture per pound of combustible. 

Other notations as in preceding equation. 

For example, the heat loss due to the moisture in a pound of fuel 
containing 10 per cent water, temperature of fuel 80 degrees F., chimney 
temperature 480 degrees F., is 

h t = 0.1 [(212 - 80) + 966* + 0.48f (480 - 212)] 
= 122.6 B.T.U. 

A rough rule is to allow a loss of 1 per cent of the total heat value of 
the dry fuel for each 10 per cent of moisture present. 

28. Loss due to the Presence of Hydrogen in the Fuel. — The hydro- 
gen in any fuel which is not rendered inert by oxygen burns to water 
and in so doing liberates 62,032 B.T.U. per pound. All of this heat is 
not available for producing steam in the boiler, since the water formed 
by combustion is discharged with the flue gases as superheated steam 
at chimney temperature. This loss is equal to 

h 5 = 9H [(212 - t) + 966* + 0.48f (T - 212)], (10) 

in which h 5 = B.T.U. lost per pound of combustible. 

H = weight of hydrogen per pound of combustible. 

All other notations as in equations (8) and (9). 

With anthracite coal this loss is approximately 2.5 per cent of the 
total heat value of the combustible and with bituminous coal it runs as 
high as 4.5 per cent. 

29. Loss due to Smoke. — Visible smoke consists of carbon in a 
flocculent state mixed with the products of combustion. It is seldom 
evident in connection with anthracite coal and is generally associated 
with bituminous fuel. A smoky chimney does not necessarily indicate 
an inefficient furnace, since the losses due to visible smoke generation 
seldom exceed 2 per cent; as a matter of fact, a smoky chimney may 
be much more economical than one which is smokeless. That is to 
say, a furnace operating with minimum air supply may cause dense 
clouds of smoke and still give a higher evaporation than one made 
smokeless by a very large excess of air. There will be some loss due 
to carbon monoxide and unburned carbon or soot in the former case, 
but this may be more than offset by the excessive losses caused by the 

* See footnote, p. 88. f See footnote, p. 37. 






FUELS AND COMBUSTION 



39 



heat carried away in the chimney gases in the latter. Judging from the 
results of the majority of steam power plants using bituminous coal, 
even those recently installed, smokeless and efficient combustion is 
not readily effected and the problem is far from being satisfactorily 
solved. 

Smoke has become such a public nuisance, particularly in the larger 
cities, that special ordinances prohibiting its production have been 
enacted and violators are subject to heavy fines. Effective enforce- 
ment of these ordinances renders smoke production very costly and 
the problem of smokeless combustion becomes a momentous one. 

The subject of smoke prevention and smoke-prevention devices is 
discussed at some length in Chapter V. 

30. Radiation and .Minor Losses. — These losses are usually deter- 
mined by difference. That is, the difference between the heat repre- 
sented in the steam and the losses just mentioned is charged to radia- 
tion, leakage, and unaccounted for. Summing up the various losses 
we have 



Heat given to steam 

Loss in chimney gases 

Loss due to carbon burning to CO 

Loss of fuel through grate 

Loss due to moisture in coal, moisture 

in air, and hydrogen in fuel 

Smoke, soot, etc 

Radiation and minor losses 



Excellent 


Good 


Average 


Practice. 


Practice. 


Practice. 


Per Cent. 


Per Cent. 


Per Cent. 


80 


70 


60 


12 


18 


24 





1 


2 


0.5 


1 


2 


3.0 


3 


3 





0.5 


1 


4.5 


6.5 


8 



Poor 
Practice. 
Per Cent. 



50 

30 

3 

3 



3.5 
1.5 

9 



31. Size of Coal. — Bituminous. Coal is usually marketed in 
different sizes, ranging from lump coal to screenings. The latter furnish 
by far the greater part of the stoker fuel used. For maximum efficiency 
coal should be uniform in size. With hand-fired furnaces there is 
usually no limit to its fineness and larger sizes can be used than with 
stokers. As a rule the percentage of ash increases as the size of coal 
decreases. This is due to the fact that all of the fine foreign matter 
separated from larger coal, or which comes from roof or floor of the 
mine, naturally finds its way into the smaller coal. The size best adapted 
for a given case is dependent upon the intensity of draft, kind of stoker 
or grate, and the method of firing, and its proper selection often affords 
an opportunity to effect considerable economy. Some idea of the 
influence of the size of screenings on the capacity and efficiency of a 



40 



STEAM POWER PLANT ENGINEERING 



boiler in a specific case is illustrated in Fig. 6. The curves are plotted 
from a series of tests conducted with Illinois screenings on a 500-horse- 
power B. & W. boiler, equipped with chain grates, at the power house 
of the Chicago Edison Company. 

Influence" of Thickness of Fire. — See paragraph 76. 

Size of Coal: Some Characteristics of Coal as affecting Performances with Steam 
Boilers : Jour. West. Soc. Engrs., Oct., 1906, p. 528. Small Size Anthracite : Eng. and 
Min. Jour., Dec. 22, 1904. The Economy of Small Size Coals for Power Plants : Eng. 
Mag., Feb., 1905. 



1000 




1.25 



1.00 



0.75 
Size of Coal in Inches 



0.50 



0.25 



Fig. 6. Influence of Size of Coal on Boiler Capacity and Efficiency. 

32. Washed Coal. — The washing of coal is for the purpose of 
separating from it such impurities as slate, sulphur, bone coal, and ash. 
All of these impurities show themselves in the ash when the coal is 
burned. Screenings contain anywhere from 5 per cent to 25 per cent 



FUELS AND COMBUSTION 



41 



of ash and from 1 per cent to 4 per cent of sulphur. Washing eliminates 
about 50 per cent of the ash and some of the sulphur. Table 6 gives 
some idea of the effects of washing upon a number of grades of coal. 
The evaporative power of the combustible is practically unaffected 
by washing and the greater part of the water taken up by the coal is 
removed by thorough drainage. Many coals otherwise worthless as 
steam coals are rendered marketable by washing. Washed coals are 
usually graded as follows : 



Size. 


Screens. 


No. 1 


Over If 


Under 2£ 


2 


11 


U 


3 


1 


It 


4 


* 


I 


5 




i 



Numbers 3 and 4 are excellent sizes for use in connection with stokers 
and No. 5 is well adapted for hand furnaces where smoke prevention is 
essential. 

TABLE 6. 
EFFECT OF WASHING ON BITUMINOUS COALS. 

(Journal W.S.E., December, 1901.) 



Before Washing. 
(Per Cent.) 



Ash. 



Sul- 
phur. 



Fixed 
Carbon. 



After Washing. 
(Per Cent.) 



Ash. 



Sul- 
phur. 



Fixed 
Carbon. 



Belt Mountain, Mont 

Wellington Colliery Co., Van- 
couver Island (new coal) . . . 

Alexandria Coal Co., Crabtree, 
Pa 

DeSoto, 111 , 

Northwestern Improvement 
Co., Roslyn, Wash 

Luhrig Coal Co., Zaleski, Ohio 

Rocky Ford Coal Co., Red 
Lodge, Mont 

Buckeye Coal and Ry. Co., 
Nelsonville, Ohio 

New Ohio Washed Coal Co., 
Carterville, 111 



18.74 

35.00 

10.60 
18.00 

16.30 
15.80 

25.30 

13.77 

9.48 



3.34 



43.72 
38.00 



1.30 



0.57 
1.90 



44.00 
45.90 



1.05 
0.78 



37.80 
49.04 
55.00 



5.56 

8.90 

6.21 
4.20 

9.70 
8.00 

8.50 

4.30 

4.85 



2.40 



0.61 



0.40 
0.87 



0.89 
0.69 



48.39 
56.90 



57.00 

47.86 
50.90 

47.20 

54.82 

63.00 



42 STEAM POWER PLANT ENGINEERING 

Modern Method of Coal Washing : Eng. and Min. Jour., May 9, 1903, Oct. 13, 1904. 
Principles of Coal Washing : Mines and Min., Aug., 1903. Bituminous Coal Washing : 
Mines and Min., April, 1905. Washing of Bituminous Coals by Luhrig Process : Jour. 
West. Soc. Engrs., Dec., 1901. Coal Washing : Jour. Soc. Chem. Ind., April 30, 1904. 
Anthracite Washeries : Eng. and Min. Jour., April 28, 1906; Col. Guard, April 
20, 1906; Trans. Am. Inst. Min. Engrs., Nov., 1905. Studies on Coal Wash- 
ing : Col. Guard, Nov. 21, 1902. Coal Washing by Stuart System : Mines and Min., 
Dec, 1903. Coal Washing at Collinsville, III: Mines and Min., Sept., 1901. Bellevue 
Washery of D. L. and W. R.R. Co., Scranton, Pa.: Mines and Min., June, 1903. Coal 
Washery at Howe, Indian Territory : Mines and Min., March 14, 1904. Eastern Coal 
and Coke Co.'s Washery at Kansas : Eng. and Min. Jour., Sept. 20, 1902. 

33. Purchasing Coal.* — Engineers fail to agree as to the specifica- 
tions best suited for the purchase of coal. Some extensive purchasers 
require elaborate analyses and others specify only the size and grade 
of the fuel. Every essential requirement of the purchaser may be 
fulfilled by confining them to the four following characteristics : 

Moisture. 

Ash. 

Size of coal. 

Calorific value of the coal. 

Although moisture is a great and uncertain variable, and the producer 
can exercise no control over this factor, still the purchaser should pro- 
tect himself against excessive moisture by stipulating an amount con- 
sistent with the average inherent moisture in the coal, and proper penalty 
should be fixed for delivery in excess of the amount allowed, a corre- 
sponding bonus being paid for delivery of less than contract amount. 
Considerable attention should be given to the percentage of earthy 
matter contained. The amount of earthy matter usually fixes the 
heating value of the coal, since the heating value of the combustible 
is practically constant. The effect of ash on the heat value of Illinois 
screenings as fired under a B. & W. boiler with chain grate is shown 
in Fig. 7. This value varies with the different types of boilers, grates, 
and furnaces, but is substantially as illustrated. The amount of refuse 
in the ash pit is always in excess of the earthy matter as reported by 
analysis. 

The maximum allowable amount of sulphur is sometimes specified, 
since some grades of coal high in sulphur cause considerable clinker- 
ing. But sulphur is not always an indication of a clinker -producing 
ash, and a more rational procedure would be to classify a coal as 
clinkering or non-clinkering according to its behavior in the particular 
furnace in question, irrespective of the amount of sulphur present. 
An analysis of the various constituents of the ash is necessary to 

* See also Selection of Coal for Boiler Furnaces, by D. T. Randall, Power & Engr., 
Apr. 6, 1909, p. 642. 






FUELS AND COMBUSTION 



43 



determine whether or not the sulphur unites with them to produce a 
fusible slag, and as such analyses are usually out of the question on 
account of the expense attached, they may well be omitted. 

The heating value of the coal as determined by a sample burned in an 
atmosphere of oxygen does not give its evaporative power, since this 

100 



90 



80 



TO 



g50 

a 
8 

40 



•20 



10 





































































































































































































































































































































































































































































































































Influence of Ash on Fuel Value of Dry 
Coal. (Illinois Screenings) 
B.& W. Boiler, Chain Grate. 
Screenings with 12.5 Per Cent Ash 
taken at 100. 










\ 














\ 
































\ 






















Jour.S.W.E 


.Oct.l 


906 P. 


542. \ 





10 20 30 40 

Per Cent of Ash in Dry Coal 

Fig. 7. Influence of Ash on Fuel Value of Dry Coal. 

depends largely upon the composition of the fuel, character of grate, 
and conditions of operation. It merely serves as a basis upon which to 
determine the efficiency of the furnace. In large plants where a number 
of grades of fuel are available it is customary to conduct a series of 
tests with the different grades and sizes, and the one which evaporates 



44 STEAM POWER PLANT ENGINEERING 

the most water for a given sum of money, other conditions permitting, 
is the one usually contracted for. In designing a new plant particular 
attention should be paid to the performance of similar plants already 
in operation, and that fuel and stoker should be selected which are found 
to give the best returns for the money. Where smoke prevention is a 
necessity the smoke factor greatly influences the choice of fuel and 
stoker. 

See paragraph 416. * 

Testing and Purchasing Coal for Steam Plants : Eng. News, Feb. 7, 1907; Eng. Rec, 
Sept. 22, 1906, p. 326; Engr. U.S., Aug. 15, 1907; Bulletin No. 339 U.S. Geological 
Survey, 1908. Coal for Hand-Fired Furnaces: Nat. Engr., July, 1909. 

34. Powdered Coal. — The value of powdered coal as a fuel for steam 
boiler plants has long been known, and appliances for pulverizing and 
feeding the coal have been on the market for a number of years. How- 
ever, despite the many advantages of powdered fuel and the apparent 
success of some of the systems of burning it, little progress has been 
made toward its general adoption. 

Some of the advantages obtained in burning powdered coal are: 

a. Complete combustion and total absence of smoke. The coal 
in the form of dry impalpable dust is induced or forced into the zone 
of combustion, where each minute particle is brought into contact 
with the necessary amount of air and complete oxidation is effected 
without the excess of air which accompanies the firing with lump coal, 
provided the furnace is properly proportioned. With a properly 
designed setting there is complete absence of smoke. 

b. A cheaper grade of bituminous coal may be burned, since the 
per cent of ash and moisture has little effect on the completeness of 
combustion and the full value of the fuel is more nearly realized than 
with ordinary firing. 

c. The plant may be rapidly forced above its rated capacity and 
sudden demands for power readily met. 

d. The labor of firing is reduced to a minimum. 

Pulverized Fuel : Eng. Mag., Jan., 1908; Jour. West. Soc. Engrs., Feb., 1904; Am. 
Elecn., Sept., 1901. Coal Dust for Steam Making : Engr. U.S., Feb. 15, 1899. Burn- 
ing Pulverized Coal: Eng. and Min. Jour., Dec. 31, 1903, May 12, 1906; Jour. Assn. 
Eng. Soc., July, 1903. Use of Pulverized Coal under Steam Boilers: Eng. News, 
April 1, 1904; Power, March, 1904, April, 1904. Coal Dust Firing: Eng. and Min. 
Jour., Dec. 16, 1905. Coal Dust Fuel: Engr., Lond., Jan. 31, 1896; Engr. U.S., 
April 1, 1903; Eng. News, Feb. 20, 1902. 

35. Depreciation of Powdered-Coal Furnaces. — To withstand the 
intense heat of combustion, brickwork of the highest quality is essential, 
since common fire brick are soon reduced to a liquid slag. A good quality 



FUELS AND COMBUSTION 45 

of fire brick will withstand the heat for several months without renewals 
provided the furnace is properly enclosed, otherwise the strain of 
expansion and contraction due to alternate heating and cooling will 
crack the brick. Excellent results have been obtained from the use of 
bricks composed chiefly of the refuse from a carborundum slag, but the 
high cost has prevented their general use. 

36. Storing Powdered Fuel. — Most cities limit the storage of pow- 
dered coal to such a small quantity as to prohibit the use of fuel feeders 
of the " dust feed " type in plants of any size not provided with a 
pulverizing and crushing system. Coal dust mixed with air is often 
claimed to be of an explosive nature and many accidents are reported 
to have resulted from this cause. Many engineers, however, refute 
this on the basis of experiments which show that explosion can only 
occur at temperatures high enough to drive off the volatile gases.,* 

37. Rate of Combustion with Powdered Fuel. — In forcing large quan- 
tities of dust into the furnace the velocity imparted to the particles may 
be so great as to carry them beyond the zone of combustion before oxida- 
tion is complete, with the result that the flues, and the back of the fur- 
nace, become covered with unconsumed carbon. So much depends upon 
the depth of the furnace and the arrangement of the regenerative surface 
that no specific figures can be given as to the maximum rate of combus- 
tion that can be efficiently effected. At ordinary rates of combustion the 
small particles of fuel are completely oxidized while in the combustion 
chamber and there is total absence of smoke. The use of anthracite 
coal is practically excluded from this type of stoker unless mixed with 
coal high in volatile matter. This is due to the fact that fixed carbon 
burns more slowly than the hydrocarbon gases and the temperature of 
ignition is higher, hence the most gentle draft will carry away the 
particles before they are completely consumed. With fuels high in vola- 
tile matter the hydrocarbons are distilled at a comparatively low tem- 
perature, forming an inflammable gas which burns rapidly with the fixed 
carbon. A mixture of 30 per cent bituminous and 70 per cent anthra- 
cite has been successfully burned in the powdered form. 

38. Cost of Pulverizing Coal. — In stokers of the " Aero Pulverizer " 
type in which the grinding and feeding are carried on simultaneously 
in a self-contained apparatus, the power consumed varies from 2 per 
cent to 10 per cent of the total power developed, depending upon the 
nature of the fuel, the load factor, the efficiency of the driving mechan- 
ism, and the degree of fineness of the powdered fuel; 5 per cent is a 
fair average. The best results are obtained when 95 per cent of the 
dust will pass a 100 mesh and 75 per cent a 200 mesh, though satis- 
factory results have been obtained with as low as 40 mesh. Powdered 

* See Fuel, Jan. 12, 1909, p. 294. 



46 



STEAM POWER PLANT ENGINEERING 



coal in the open market ranges from 25 cents to 50 cents a ton above 
the price of the same coal in the form of screenings. 

39. Efficiency of Powdered-Coal Furnaces. — Table 7 gives the 
results of a comparative test of a 140-horse-power Babcock & Wilcox 
boiler, hand fired, vs. coal-dust feeder. The test was conducted by 
the engineering staff of the McCormick Harvester Company at Chicago, 
Illinois, and the results were obtained with boilers working under normal 
conditions. The dust apparatus was a modified " Ruhl " feeder, and 
was installed by the C. 0. Bartlett & Snow Company of Cleveland, Ohio. 
In this particular test the efficiency of the boiler is very low for both 
hand fired and dust-feed, but the dust-feed test shows an efficiency of 
10 per cent greater than that of the hand fired. 

TABLE 7. 

COMPARATIVE TEST OF 140-HORSE-POWER BABCOCK & WILCOX BOILER. 

Hand Fired vs. Pulverized Fuel. 



Boilers fired by 

Date 

Duration of test 

Total water evap. into dry steam from 

and at 212 degrees 

Average gauge pressure 

Average feed-water temperature, Fahr. . . . 

Average stack-gas temperature, Fahr 

Kind of coal used 

Cost of coal delivered in boiler room ready 
to fire 

Total weight of dry coal consumed 

Per cent of ash in coal determined by lab- 
oratory analysis 

Per cent of ash as removed from ash pit 
and furnace 

Heating value of coal 

Water evap. per pound of fuel, actual con- 
ditions 

Equivalent water evap. from and at 212 
degrees per pound of dry fuel 

Equivalent water evap. per pound com- 
bustible 

Horse power developed 

Dry fuel per hour per square foot grate 
surface 

Equivalent water evap. per hour per 
square foot heating surface 

Cost per 1,000 pounds water evaporated (for 
fuel ready to fire only) 

Efficiency of boiler and furnace based on 
coal ; 



machine 
2-24r-04 
8 hours 


hand 
2-3-04 
8 hours 


70,070 

79.8 1b. 

169.7 

506 

Westville, Indi- 


45,673 

79.4 1b. 

172.1 

458 

Westville 


ana, screenings 
pulverized to 
40 mesh 


screenings. 


$2.10 
9,373 


$1.72 
8,413 


17.5 


19.54 


none 
12,555 


20.57 
11,300 


6.822 1b. 


4.595 lb. 


7.476 1b. 


5.429 1b. 


9.132 lb. 
254 


6.941 lb. 

165.5 


19.27 


18.65 


3.128 


2.039 


$0.1455 


$0,177 


55.5 per cent 


41.5 per cen 



FUELS AND COMBUSTION 47 

A comparison of a number of tests of hand-fired and powdered-coal 
furnaces with different types of feeders shows a decided gain in efficiency 
of the powdered coal over the hand-fired where the fuel is of a low grade. 
The gain becomes less marked with fuel of fair quality and disappears 
entirely with good fuel and properly manipulated automatic stokers. 
A test made by G. H. Barrus on a 250-horse-power B. &. W. boiler at 
the General Electric Works in connection with a coal-dust feeder manu- 
factured by the Phcenix Investment Company of New York gave a boiler 
and furnace efficiency of 75.3 per cent. Subtracting from this the power 
consumption of 5 per cent for operating the crusher and feeder, the net 
efficiency was 70.1 per cent. A test of a 135-horse-power return tubular 
boiler with this same stoker gave a combined efficiency of boiler and 
furnace of 80 per cent. These figures, however, have been equaled and 
even exceeded in special hand-fired automatic stoker tests, and only a 
comparative test of the two systems under similar conditions will show 
their respective efficiencies. 

Tests of Pulverized Fuel: Engr. U.S., April 1, 1904; Engr. Lond., Jan. 31, May, 
1904; Power, May, 1904. Comparative Boiler Tests with Ordinary and Pulverized 
Coal Firing : Eng. Rec, March 12, 1904. 

40. Furnaces for Burning Powdered Coal. — In burning ordinary bulk 
coal the mass of incandescent fuel stores up a sufficient quantity of heat 
to effect the distillation and ignition of the volatile matter in the green 
fuel. With pulverized coal a refractory lining is necessary to bring 
about the same result. In arranging a furnace for burning powdered 
coal in connection with a burner of the forced draft type, the grate bars 
are removed, ash and fire doors bricked up, and the nozzle bricked in 
tightly. The lower surfaces of the tubes are covered and the whole 
forms a reverberatory furnace. With the natural draft system of 
burner, a suitable opening is left in the brick lining of the ash door to 
allow the necessary amount of air for combustion to enter. Considerable 
difficulty is found with delivery nozzles in the formation of slag in the 
outlet and in their rapid destruction on account of the intense heat. A 
water-jacketed cast-iron nozzle is said to satisfactorily overcome these 
objections. 

41. Draft for Powdered Fuel. — A study of a number of tests of 
boilers burning powdered coal shows that the necessary draft is very 
low and ranges from 0.05 to 0.2 of an inch of water and averages not far 
from 0.1 inch. 

42. Types of Powdered-Coal Burners. — Powdered-coal burners may 
be grouped into two general classes: , 

1. The dust-feed burner, in which the coal is supplied in the powdered 
form, and 



48 



STEAM POWER PLANT ENGINEERING 



2. The self-contained burner, in which the coal is crushed, pulver- 
ized, and fed to the furnace simultaneously. 
The dust may be fed into the furnace by 

1. Natural draft. 

2. Mechanical means, or by 

3. Forced draft. 

The following outline gives a classification of a few of the best known 
coal-dust burners : 



Natural Draft 



Forced Draft 



Natural Draft 
Feed 

Brush Feed 



Blower Feed 



fpinther 
j Wegener 

Schwartzkopff 

/Cyclone 
"(Triumph 



Dust Feed 



Compressed Air /Eng and Powdered 
(Fuel Company 



Paddle Wheel 



/Ideal 

(.Aero-Pulverizer 



Self-contained 



43. Pinther Apparatus. — Fig. 8 shows a section through a Pinther 
coal-dust feeder, illustrating the principles of the " natural draft feed " 

type. The powdered coal is 
placed into hopper B, from 
which it is fed by rollers a, a 
into the chamber leading to 
the furnace C. The dust falls 
in a thin stream and is caught 
up by the current of air and 
drawn into the furnace as 
indicated. The furnace is 
lined with refractory material 
heated to a sufficiently high 
temperature to ignite the fuel 
and burn it in suspension. The 
chief drawback to a burner 
of this type is its limited 
capacity. Any attempt to 
feed large quantities of fuel 
into the furnace necessitates 

Fig. 8. Pinther Coal-dust Feeder. guc ] 1 & s t r0n g current of air as 

to carry the particles of dust beyond the zone of combustion before they 
are completely consumed. Within the limits of its capacity it is an 
efficient and simple apparatus, but is open to the same objection as all 




FURNACE 



FUELS AND COMBUSTION 



49 



FURNACE 



burners of this type in that it necessitates the storage of powdered 
coal. This apparatus is not much in evidence in boiler plants. 

44. Schwartzkopff Apparatus. — Fig. 9 shows a section through a 
Schwartz kopff feeder, illustrating 
the principles of the brush-feed, 
naturaWraft system. It is a 
very simple and practical dust 
feeder, though open to the objec- 
tion of all systems which require 
the coal to be ground and pulver- 
ized in separate machines. The 
fuel is placed in a hopper and its 
supply to the brush is regulated 
by the hand screw A and the 
spring plate bottom of the 
hopper. The brush, consisting 
of a number of flat steel leaves 
^ inch by J inch wide, revolves 
at a high speed, 1000 to 1200 
r.p.m. and forces the dust into 
the furnace. The air for combus- 
tion is admitted either through 
the grates in the ordinary way or 
through the lower chamber of 
the burner. To prevent the dust 

from bridging in the hopper, a small hammer C is fitted to the brush so 
that it will strike the plate D and agitate the dust. This apparatus is 
meeting with much success in connection with annealing furnaces, but 
is still in the experimental state as far as boiler firing is concerned. 

45. Aero-Pulverizer Apparatus. — Fig. 10 gives a general view of the 
Aero-Pulverizer Company's apparatus, and is a typical example of a self- 
contained system. It is very compact, occupying a floor space of only 
30 by 77 inches, and is capable of burning 300 to 1500 pounds of coal per 
hour. It consists essentially of four interior communicating chambers 
of successively increased diameter in which paddles revolve on. arms 
with corresponding increased radii. The largest chamber contains a 
fan, the function of which is to draw the pulverized material successively 
from one chamber to another and to finally deliver it through the exit 
in the fan chamber under the impetus of a forced draft. There are two 
adjustable inlets for air at the feed of the machine through which is 
introduced the amount of air required for pulverizing purposes. The 
apparatus may be belt driven or direct connected and runs at about 




Fig. 9. Schwartzkopff Coal-dust Feeder. 



50 



STEAM POWER PLANT ENGINEERING 



1600 r.p.m., requiring from 6 to 15 horse power for its operation. It is a 
complete dust fuel feeding system on one bed plate comprising a pul- 
verizer, fan, coal feeder, hopper, and air dampers. The operation is as 
follows: Coal previously crushed to nut size is fed to the hopper from 



TOP CASING THROWN BACK 
FOR INSPECTIO 



COAL 
HOPPER 



AUTOMATIC 
FEEDER 




Fig. 10. Aero-Pulverizer Coal-dust Feeder. 

the bottom of which it is transferred, with the necessary air for com- 
bustion, to the pulverizer chamber. The coal, passing into the pul- 
verizer, is thrown out radially by centrifugal force, due to the rapidly 
revolving arms and bats, and is reduced to a dust by percussion and 
attrition. The dust is moved over the ends of the bats and into the fan 
chamber from which it is blown into the furnace. This apparatus will 
successfully pulverize and feed coal containing as much as 10 per cent 
moisture. 




Fig. 11. Triumph Coal-dust Feeder. 

46. Triumph Apparatus. — Fig. 11 illustrates the Triumph coal-dust 
feeder as designed by the C. O. Bartlett & Snow Company, Cleve- 
land, Ohio. 






FUELS AND COMBUSTION 51 

The coal is fed from storage bin to hopper A and feed worm B. 
The latter forces it down spout F directly to delivery tube D, where it 
is caught by the air draft and fed into the furnace. 

The amount of feed depends upon the speed of the feed worm, which 
is driven by the friction disk / against the flange plate H. This disk is 
moved in or out by handle so as to get any speed desired. The air is 
furnished by fan C, the amount being controlled by valve E. 

DESCRIPTION OF COAL-DUST BURNERS. 

Aero-Pulverizer System: Eng. News, Nov. 28, 1901, p. 415; Eng. Rec, May 25, 

1901, p. 506; Power, March, 1904. 

Cyclone System: Engr. U.S., April 1, 1903, p. 272; Eng. News, Nov. 28, 1901, 
p. 415; Power, March, 1904. 

Davis Pulverizer : Jour. Asso. Eng. Soc, July, 1903; Engr. U.S., April 1, 1903. 

Ideal: Am. Elecn., April, 1902, p. 196; Power, March, 1904. 

Miscellaneous Coal Dust Burners: Am. Elecn., Sept., 1901, p. 434; Engr., Lond., 
Sept. 11, 1896; Engng., Jan. 15, 1897; Power, Aug., 1903; St. Ry. Rev., Vol. 8-187, 
1898; Engr. U.S., April 1, 1904. 

Rowe : Engr. U.S., Jan. 1, 1903, p. 93; April 1, 1903, p. 272; Eng. News, Dec. 25, 

1902, p. 548; Eng. Rec, Dec. 20, 1902, p. 591. 

Schwartzkopff : Am. Elecn., Jan., 1902; Eng. News, Feb. 20, 1902; Power, March, 
1904. 

Wegener : Cassier's, March, 1896, p. 501; Power, March, 1904; Eng. Mag., March, 
1896, p. 1158, Aug., 1896, p. 964, Oct., 1898, p. 125; Eng. News, Sept. 16, 1897, 
p. 189, 

47. Fuel Oil. — The recent development of oil wells in the Western 
and Gulf States, with the consequent enormous increase in production, 
has given a marked impulse to the use of crude oil for fuel purposes in 
steam power plants. Where economic and commercial conditions 
permit, it is the most desirable substitute for coal. The total absence 
of smoke and ashes, prompt kindling and extinguishing of fires, extreme 
rate of combustion, and ease with which it can be handled and con- 
trolled are marked advantages in favor of fuel oil. The reduction in 
volume and weight over an equivalent quantity of coal for equal heat- 
ing values and the increase in boiler efficiency are factors of no mean 
importance, particularly in connection with marine or locomotive work. 
In stationary work the chief objections are the difficulty in securing 
ample storage capacity and the increased rate of insurance. An objec- 
tion sometimes raised against oil fuel is the increased depreciation of 
the setting, but in a well-designed setting this figure is only nominal and 
of secondary importance. However, in spite of the many advantages 
presented in the use of fuel oil for power plant purposes, the limited 
supply and constant fluctuation in price prevent its adoption as a 
general fuel, and limit its use to the plants most favorably located. 



52 



STEAM POWER PLANT ENGINEERING 



Crude Oil Burning : Power, March, 1907; Engr. U.S., Dec. 15, 1905, March 1, 1906; 
Am. Elecn., Aug., 1903, p. 396; Engng., March 28, 1902, p. 140; Eng. Mag., May, 
July, Sept., 1903; Eng. News, June 19, 1902, p. 501; Cassier's, May, 1901, p. 61; 
Engr., Lond., Dec. 9, 1904; St. Ry. Jour., May 10, 1902, p. 588; Am. Gas Light 
Jour., May 12, 1902, p. 695, 



. TABLE 8. 
ANALYSES OF TYPICAL AMERICAN FUEL OILS. 





Authority. 


Physical Properties. 


Location. 


Specific 

Gravity. 

60°-70° F. 


Flash 
Point. 
Deg. F. 


Burn- 
ing 
Point. 
Deg. F. 


Specific 
Viscosity. 




60° F. 


185° F. 


California — Crude 


Ed. O'Neill 

....do 


0.9533 
0.9572. 






299.6 
373 
1.17 


4.7 


Do 








Do 


do 


0.7825 
0.9670 
0.866 


62 

196 

52 


64.5 
221 

77 




Do 


....do 




Kansas — Crude 


B. F. McFarland. 

C. E. Coates 






Louisiana — Crude 






Ohio — Distillate 


Deville 


0.887 
0.838 
0.826 
0.886 
0.841 










Do 


N. W. Lord 

Deville 


177 


212 






Pennsylvania — Crude . . 
Pennsylvania — Distillate 
W. Virginia — Crude. . . . 

Wyoming — Crude 

Texas — Crude 






do 










..do 










Colburn 










Denton 


0.92 
0.926 


142 
216 


181 
240 






Texas — Distillate 


U. S. Naval Re- 
port 













Location. 


Authority. 


Chemical Properties. 


C 


H 


O+N 


S 


B.T.U. 
per Lb. 


California — Crude 

Do 


Ed. O'Neill 

do 


85.75 
86.3 


11.3 
10.7 




0.668 
0.8 


18,797 
18,646 


Do 


... .do 


Do 


.. ..do 












Kansas — Crude 


B. F. McFarland. 

C. E. Coates 


85.4 


13.07 








Louisiana — Crude 




0.34 


19,814 
18,718 
19,880 
17,930 
19,210 
18,400 
19,590 
19,060 


Ohio — Distillate 


Deville 


84.2 


13.1 


2.7 


Do 


N. W. Lord 


Pennsylvania — Crude . . 
Pennsylvania — Distillate 
W Virginia — Crude .... 


Deville 


82 

84.9 

84.3 


14.8 
13.7 
14.1 


3.2 
1.4 
1.6 




....do 


....do 


Wyoming — Crude 

Texas — Crude 




Denton 


84.6 
83.26 


10.9 
12.41 


2.87 
3.83 


1.63 
0.50 


Texas — Distillate 


U. S. Naval Re- 
port 


19,481 



48. Chemical and Physical Properties of Fuel Oil. — From Table 8 
it will be seen that the physical properties of oils from different localities 
in the United States differ widely, while the chemical constituents vary 



FUELS AND COMBUSTION 



53 



but slightly. For example, the oils given in the table differ greatly in 
volatility, specific gravity, and viscosity, but have nearly a constant 
ratio of carbon and hydrogen and consequently vary but slightly in 
heating value. 

A good deal of the oil produced is unfit for fuel purposes unless 
refined. The chief impurities are sulphur, earthy matter, and water. 
Besides interfering with the free burning of the oil, moisture and sulphur 
have a deleterious effect upon the boiler and furnaces, and should not 
be present in large quantities. Where the percentage of sulphur is 
greater than 4 per cent, the depreciation of the boiler and furnace 
offsets the gain in using the lower grade fuel. Many successful processes 
of removing the water and sulphur are on the market, and consequently 
crude oil high in sulphur should not be used unless the depreciation 
element has been taken into consideration. 

Oil that is to be transported or stored or used for fuel inside of 
buildings should be of the " reduced " variety, from which the naphtha 
and higher illuminating products have been distilled. The gravities 
of such distillates vary from 20 to 25 degrees Baume, or close to 0.9 
specific gravity, and their flash points range from 240 degrees. F. to 270 
degrees F. This variation in volatility has little effect on the heat 
value of the oil, since the ratio of carbon to hydrogen varies but slightly 
in the various distillates. 

One barrel of oil contains 42 gallons and weighs from 310 to 332 
pounds according to the specific gravity. Compared with coal, oil 
occupies about 50 per cent less space and is 35 per cent less in weight 
for equal heat values. The comparative heat values of coal and oil 
are approximately as follows: 



B.T.U. per Pound 
of Coal. 


Pounds of Coal Equal 
to 1 Barrel of Oil. 


Barrels of Oil Equal 

to 1 Short Ton of 

Coal. 


10,000 
11,000 
12,000 
13,000 
14,000 
15,000 


620 

564 
517 
477 
443 
413 


3.23 
3.55 

3.87 
4.19 
4.52 

4.84 



Technical Aspects of Oil as Fuel: Junge, Power, Oct., 1907, p. 665. 

Petroleum Oil Fields : Jour. Soc. Chem. Ind., Oct. 15, 1902, p. 1228. Investiga- 
tion on American Petroleum : Am. Chem. Jour., March, 1896, p. 215. The Constit- 
uents of Pennsylvania, Ohio, and Canada Petroleums: Amer. Chem. Jour., Vol. 19-419. 
Composition of California Petroleum : Amer. Chem. Jour., Vol. 19-796, Vol. 25-253. 
Composition of Petroleum: Amer. Chem. Jour., Vol. 28-165, 33-251. Composition 



54 



STEAM POWER PLANT ENGINEERING 



of Texas Petroleum: Jour. Amer. Chem. Soc, Feb. 9, 1901, p. 264; Soc. Chem. Ind., 
19-121, 20-237, 690. Flashing Points of Petroleum : Jour. Soc. Chem. Ind., 15-341. 
Origin of Petroleum : Jour. Soc. Chem. Ind., 16-727, 1898. Influence of Water on 
Flash Test and Combustion Point of Petroleum : Chem. News, 85-267. The Relation 
between Some Physical Properties of Bitumens and Oils : Eng. Rec, Aug. 18, 1906. 

49. Efficiency of Boilers with Fuel Oil. — From Table 14, it will be 
seen that 70 per cent is a high figure for boiler efficiency in regular 
service when good coal is burned, and 65 per cent a fair average. With 
liquid fuel an average efficiency of 4 to 6 per cent above this is readily 
attained. (See Table 9.) This increase in efficiency is partly due to 
the fact that the oil is readily broken up and brought into intimate 
contact with the necessary air for combustion and the loss due to excess 
of air is correspondingly reduced. The results of tests made by the 
Liquid Fuel Board of the U.S. Navy show that oil has an efficiency of 
5 per cent greater than coal for the same rate of evaporation, and that 
the boiler capacity may be increased 50 per cent above that possible 
when burning coal and still maintain the same efficiency. The max- 
imum efficiency with oil was attained at a higher rate of evaporation 
than was the maximum efficiency when coal was burned. 

TABLE 9. 

BOILER EFFICIENCIES, OIL FUEL. 



X 
0) 

a 
i— ( 


Authority. 


Reference. 


Quality of Oil. 


Evapora- 
tion from 
and at 
212° F. 
Pounds. 


Efficiency 
of Boiler 
and Fur- 
nace. 
Per Cent. 


1 
2 


Pacific Light and 
Power Co., Los 
Angeles, Cal. 


Eng. Record, Aug. 
6, 1904. 


California Crude, 
18,607 B.T.U. per 

pound. 
California Crude, 

18,760 B.T.U. 


16.02 
15.66 


83.06 
80.6 


3 
4 


U.S. Naval Board 


1902 Report of 
U.S. Naval 
Liquid Fuel 
Board. 


Reduced Beau- 
mont, 19,480 
B.T.U. 

20,000 B.T.U 


14.43 
16.9 


71.5 
77.8 


5 
6 


Prof. Williston. . . 


Engineering Mag- 
azine, July, 1903. 


West Va. Crude, 
20,960 B.T.U. 

Texas Crude, 
18,850 B.T.U. 


16.5 
15.9 


76 

76.8 


7 


Prof. Denton. . . . 


Power, Feb., 1902. 


Beaumont Texas, 
19,060 B.T.U. 


15.5 


78.5 


8 


Wallsend 


Engineering, Nov. 
6, 1902. 


Not stated 


14.45 









FUELS AND COMBUSTION 55 

50. Comparative Evaporative Economy of Oil and Coal. — In deter- 
mining the comparative economy of coal and oil, the fixed and operating 
charges must be considered in addition to the cost and efficiency of 
the fuel. From the market quotation on oil and coal and the com- 
parative heating values of each the actual cost per B.T.U. is readily 
obtained, and by combining this with the relative efficiencies from the 
furnace standpoint the net cost of the fuel is obtained. The fixed 
charges vary with the location and size of the plant and are approxi- 
mately the same per boiler horse power for a given location in both 
cases. The insurance rates may be greater with the oil fuel and the 
depreciation of the boiler setting may be somewhat larger, but in a well- 
constructed furnace the latter item should be the same in both instances 
for average rates of combustion. The operating charges are decidedly 
in favor of the oil fuel, since no ash handling is necessary. Oil fuel is 
readily fed to the furnace, and the cost of attendance may be materially 
less than with coal firing, and one man may safely control from eight to 
ten boilers. Table 106, Chapter XVII, gives data relative to the cost 
of producing electrical power in connection with oil-fired steam plants. 

Tests of Crude Oil as a Fuel : Cassier's, May, 1901, p. 61 ; Power, Feb., 1902, p. 8; 
Eng. U.S., Jan. 16, 1905, p. 90, Feb. 1, 1905; Eng. Rec., Dec. 20, 1902;Eng. Mag., 
July, 1902, p. 615; Eng. News, July 11, 1901, p. 23; Eng. Rec, Aug. 6, 1904, p. 175; 
Oct. 29, 1904, p. 502. 

51. Oil Burners. — The function of the burner is to atomize the oil 
to as nearly a gaseous state as possible. 

Classification of a few well-known burners: 

Mechanical Spray : 

Korting. 
Vapor or Carburettor : 

Durr. 

Harvey. 

Spray Burners : 
Outside Mixers. 

a. Booth. 

b. Warren. 

Inside Mixers. 

a. Hammel. 

b. Kirkwood. 

c. Branch. 

d. Williams. 



56 



STEAM POWER PLANT ENGINEERING 




Oil burners for burning liquid fuel may be divided into three general 
classes : 

1. Mechanical spray, in which the oil previously heated to a tem- 
perature of about 150 degrees F. is forced under pressure through 
nozzles so designed as to break it up into a fine spray. The Korting 

Liquid Fuel Burner, Fig. 12, is an 
example of this type. In this 

(~\ \ ffi|yr~ ""jtffHil design a central spindle, spirally 

grooved, imparts a rotary motion 
to the oil and causes it to fly into a 
spray by centrifugal force on issuing 
from the nozzle. The particles of 
oil are burned in the furnace when 
they come in contact with the neces- 
sary air to effect combustion. This 
type of burner is little used in this 

country in connection with power plant work, but is meeting with 

much success on the continent. 

2. Vapor burners, or carburettors, in which the oil is volatilized in 
a heater or chamber and then admitted to the furnace, are seldom used 
except in connection with refined oils, as the residuals from crude oil 
are vaporized only at a high temperature. The Durr and Harvey 
gasifiers are the best known of this type. 

3. Spray burners are by far the most common in use. In this type 
the oil is held in suspension and forced into the furnace by means of a 
jet of steam or compressed air. Spray burners are designed either as 
outside mixers, in which the oil and atomizing medium meet outside the 
apparatus, or inside mixers, in which the oil and atomizing medium 
mingle inside the apparatus. 



Fig. 12. 



OIL 

Korting Fuel Oil Burner. 




OIL 



'&aaawzzz£azzz££&zuaa&4. 



izEzzzzzznzzzzzzzzzzmzzzzzzzzzzzzzzzpzz L 



STEAM 




Fig. 13. Booth Fuel Oil Burner. 



The Booth burner, Fig. 13, illustrates the principles of the " outside 
mixer " type of apparatus and is in use on the Santa Fe Railroad. In 
this type the oil flows through a thin slit and falls upon a jet of steam 






FUELS AND COMBUSTION 



57 



which atomizes it and forces it into the furnace. A feature of this 
apparatus is its simplicity of construction and freedom from clogging. 

Fig. 14 illustrates the Hammel burner as used at the power house of 
the Pacific Light and Power Company, Los Angeles, Cal. Oil enters the 





G.I. H.- 

Fig. 14. Hammel Fuel Oil Burner. 



burner under pressure and flows through opening D to the mouth of the 
burner where it is atomized by the steam jets issuing from slots G, H, 




STEAM 



5sr«»dl3 



p 



OIL 

Fig. 15. Branch Fuel Oil Burner. 

and 7. The oil is preheated to facilitate its flow through the supply 
system. Plates K-K are removable and are easily replaced when worn 
out or burned. The Hammel burner belongs to the " inside mixers." 



58 



STEAM POWER PLANT ENGINEERING 



A few well-known types of " inside mixers " are illustrated in Figs. 
14 to 16. The operation is practically the same in all of them and they 
differ only in mechanical details. 




Fig. 16. Kirkwood Fuel Oil Burner. 




Fig. 17. Williams Fuel Oil Burner. 



The Williams burner, Fig. 17, differs somewhat from the others in 
that the air supply passes through the burner and mingles with the oil 
and steam before entering the furnace. 



FUELS AND COMBUSTION 



59 



The simplest and most reliable burners are of the Hammel type and 
are much in evidence in the Pacific States. 

Notes on Oil Burners using Compressed Air : Power, Nov., 1904. Report of U.S. 
Naval Liquid Fuel Board : Engr. U.S., Dec. 1, 1904. Oil Burners : Engng., April 15, 
1904; Am. Engr. and R.R. Jour., Sept., 1901. 




• J40ILPIPE 



% STEAM PIPE 





OIL-* 



Fig. 18. Warren Fuel Oil Burner. 



52. Furnaces for Burning Oil Fuel. — Fig. 19 shows the construction 
of a typical oil-burning furnace as applied to a 250-horse-power B. & W. 
water tube boiler in the power plant of the Union Loop Elevated Station, 
Chicago. For the successful burning of oil the furnace should be so 
constructed that oxidation of the fuel is complete before it reaches the 
tubes. This is effected by arranging the refractory lining to form a 
sort of reverberatory furnace in which the atomized oil is vaporized and 
mixed with the necessary air for combustion. The air is preheated in 
its passage beneath the lower lining of the furnace and the supply is 
regulated by a suitable damper. The regulation of air, steam, and oil 
for the burner is a very delicate operation and requires considerable 
skill for efficient results.* In the particular furnace illustrated in Fig. 19 
the flame impinges against a cellular wall of fire brick before it reaches 

* For a modern and highly efficient system of oil fuel feeding and regulation, see 
Power, Dec. 29, 1908, p. 1108. 



60 



STEAM POWER PLANT ENGINEERING 



the bridge wall. The bricks are loosely stacked and are readily removed 
when burned out. They tend to save the lining of the bridge wall arch 

and insure a more intimate 
mixture of air and oil in the 
combustion chamber. The 
supply of crude oil is limited 
compared with that of coal 
and the price is subject to 
sharp fluctuations, and conse- 
quently the cost may prove 
prohibitive for fuel purposes. 
To be prepared for such an 
emergency, many engineers 
design the furnaces for coal 
burning and arrange them 
with loose brick-work for the 
temporary use of oil. 

Fig. 20 illustrates the 
application of a Hammel 
burner at the rear end of a 
furnace. 



Oil~Fired Furnaces : Engr. U.S., 
July 1, 1902, p. 491, Nov. 15, 1905; 
Eng. Mag., May, 1903; Engng., 
April 15, 1904, p. 523, April 29, 
1904, p. 594. Liquid Fuel Com- 
bustion: Trans. A.S.M.E., May, 
1902. Equipment for Oil Fuel: 
Eng. and Min. Jour., Oct. 7, 
1905. 

53. Air vs. Steam as an 
Atomizing Medium. — Table 
10 gives the results of a 
series of tests made by the 
U.S. Naval Liquid Fuel Board 
in 1902 on different types of 
burners using air, steam, or 
both for atomizing the fuel. 
The first eight tests were 
made with compressed air 
as the spraying medium and under pressures varying from 0.78 pounds 
to 4.68 pounds per square inch. The most economical results were 




°5 



FUELS AND COMBUSTION 



61 



obtained with the lower pressures, and the total steam used to compress 
the air varied from 1.06 per cent to 7.45 per cent of the total steam 
generated, but not all burners would work with air at this low pres- 
sure. With steam as the spraying medium the steam required to 




Fig. 20. Furnace for burning Fuel Oil, Rear Feed. 



operate the burner varied from 3.98 per cent to 5.77 per cent of 
the total generated, while the burner using both steam and air 
required 6.09 per cent to 8.54 per cent of the total. The results of 
recent tests with the latest types of burners give somewhat lower 
steam consumption than the tests conducted by the Naval Board 
and a good average is not far from 3 per cent.* Table 10 contains 
also the results of a few scattering tests made with different types of 
burners. 

In general it may be said that where a supply of low-pressure air 
is available, air is unquestionably more economical than steam as an 
atomizing medium, but in the average boiler plant the use of steam 
obviates complication and risk of interrupted service. Where it is 
necessary to use high-pressure air the economy decreases with the 
increase in pressure, since the cost of each cubic foot of compressed air 
increases rapidly with the pressure, but its ability to atomize the oil 
does not increase proportionately. 

Steam vs. Air for Liquid Fuel and Oil Burners : Am. Mach., Vol. 27, No. 51. 
* See Proceedings, A. S. M. E., Dec., 1908, p. 1698. 



62 



STEAM POWER PLANT ENGINEERING 



•fttrao 
J9d) Jsnog jo iCouaptga 



•p9}T3J9U9S UlB91g 
imoj, JO ^1190 J9d t no 
Sui^jdg ui pasri uii39ig 



QOMlOrHTttOOO^OSNt^WlHN 



000«OOONW>0>ONOOiHCOtJIOS f CO <M 



•(spunoj) 
'£ Z\Z re Pub uiojj 
uoij'BJod'BAa; luai-BAinbg 



MCOrHfNjHrtNrflCCO^lOOOCO ■««« 


OOMCO<N(N051>CTiNiOOOiO©»CO>CON 


N<MT|Hi-lTttTj(MO«MMeO'*^-*>0»0'* 



(•^ S99lg9Cl) Uinip8J\[3UI 

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\\q Sui^Bidg joj pasn 

ranip8J\[ JO 9JT1SS8JJ 



(93ni3£) ipui gjenbg J9d 

SpunOj) 9JT1SS91J UI129Jg 



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lO lO lO to «o lO lO 



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<N<N<MC^(M<M<M<M<M(MCq<M<MC^i-H t— I , — ( 



o o o o o o£9, . 



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s? 



WW 



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o °o 



H? Mo 

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iH(NC0^>O©NQ005O>-<(NC0'1<>O!ON00 



FUELS AND COMBUSTION 



63 



54. Oil Pressure. — This varies with the different types of burners 
and ranges from a few pounds to 60 pounds or more per square inch. 
The low-pressure systems are ordinarily operated under standpipe pres- 
sures as in Fig. 21, which illustrates the arrangement of apparatus as 
advocated by the International Gas and Fuel Company. A steam pump 
B draws the oil from the buried tank through pipe Z and delivers it to 
the standpipe E. Thence it flows through pipe / to the burners under a 




Fig. 21. International Gas and Fuel Company's Fuel Oil System. 



head of about 10 feet. The pump runs constantly, the surplus oil flow- 
ing back to the tank through the pipe T. The oil is heated by the 
exhaust pipe Z'. The oil pump is provided with a device D having a 
piston connected by a chain with a cock S, which automatically opens 
when the boiler is not under steam pressure, so that the standpipe will 
be emptied, the oil flowing to the storage tank. 

The high-pressure systems are invariably operated by steam pumps, 
usually in duplicate, and are so arranged that the oil pressure will be 
kept practically constant irrespective of the steam pressure. The adjust- 
ment of steam and oil is a very delicate operation, and fluctuation in the 
steam pressure disturbs the proportion of oil and steam; to prevent this 



64 



STEAM POWER PLANT ENGINEERING 



the steam pressure at the burner is reduced several pounds below that 
of the boiler by suitable reducing valves and is thereby kept at a nearly 
constant value. 

55. Oil Storage and Transportation. — Distillates or reduced oils are 
readily stored and transported, but the crude oils, on account of the 
inflammability of the highly volatile elements, offer a different problem. 
In most cities distillates may be stored in large quantities but only in 
tanks sunk below the lowest level of the surrounding territory. This 
is a protection against flooding the district with burning oil in case of 
a fire. In the country the oil is ordinarily stored in tanks above the 
ground level and at some distance from the plant. 

Fig. 22 illustrates the Hydraulic Oil Storage Company's system of 
storing oil and delivering it to the burners. The oil reservoirs are placed 



SIPHON BREAKER 
1% 




OIL TO BURNERS 
DISCHARGE TO SEWER 

FLOOR LINE 



Fig. 22. Hydraulic Oil Storage Company's Fuel Oil System. 



below grade as indicated to minimize fire risk. The operation is as 
follows: Water enters the " float box" and flows through a " three-way 
cock " to the bottom of the reservoir until all of the oil and water 
pipes are filled up to the level of the float box, when the float auto- 
matically cuts off the supply. This flooding of the entire system 



FUELS AND COMBUSTION 65 

drives out all of the air. The three-way cock is then turned to 
" discharge " and part of the water flows to the sewer. The tank- 
car or wagon is next attached to the " oil inlet " and the oil flows into 
the tank and displaces the water until the level of the " filler float " 
is reached, when the supply is automatically cut off. The inlet is so 
placed that the head of oil in the tank-car is sufficiently great to over- 
come the opposing head of water. The three-way valve is next turned 
to the first position and the head of water forces the oil to the burners. 
After the oil has been withdrawn from the storage tank the water can 
only rise to the level of the water in the float box and therefore cannot 
be fed to the furnace. The small steam pipe admits steam into the 
tank and heats the oil, thereby making it flow more freely. 

Storing Oil Fuels : Eng. News, Oct. 29, 1903, p. 396. 

Petroleum Reservoirs : Jour. Soc. Chem. Ind., Jan. 31, 1899. 

Handling Fuel at Railway Terminals : Eng. News, Sept. 25, 1902, p. 232. 

56. Conclusions of U. S. Naval Liquid Fuel Board. — After a series 
of elaborate tests it was concluded 

a. That oil can be burned in a very uniform manner. 

b. That the evaporative efficiency of nearly every kind of oil per 
pound of combustible is probably the same. While the crude oil may 
be rich in hydrocarbons, it also contains sulphur, so that, after refining, 
the distilled oil has probably the same calorific value as the crude 
product. 

c. That a marine steam generator can be forced to even as high a 
degree with oil as with coal. 

d. That up to the present time no ill effects have been shown upon 
the boiler. 

e. That the firemen are disposed to favor oil, and therefore no 
impediment will be met in this respect. 

/. That the air requisite for combustion should be heated if possible 
before entering the furnace. Such action undoubtedly assists the 
gasification of the oil product. 

g. That the oil should be heated, so that it could be atomized more 
readily. 

h. That when using steam higher pressures are undoubtedly more 
advantageous than lower pressures for atomizing the oil. 

i. That under heavy forced draft conditions, and particularly when 
steam is used, the Board has not yet found it possible to prevent 
smoke from issuing from the stack, although all connected with the tests 



66 STEAM POWER PLANT ENGINEERING 

made special efforts to secure complete combustion. Particularly for 
naval purposes, it is desirable that the smoke nuisance be eradicated in 
order that the presence of a war ship might not be detected from this 
cause. As there has been a tendency of late to force the boilers of 
industrial plants, the inability to prevent the smoke nuisance under 
forced-draft conditions may have an important influence upon the 
increased use of liquid fuel. 

/. That the consumption of liquid fuel cannot probably be forced 
to as great an extent with steam as the atomizing agent as when com- 
pressed air is used for this purpose. This is probably due to the fact 
that the air used for atomizing purposes, after entering the furnace, 
supplies oxygen for the combustible, while in the case of steam 
the rarefied vapor simply displaces air that is needed to complete 
combustion. 

k. That the efficiency of oil-fuel plants will be greatly dependent 
upon the general character of the installation of auxiliaries and fittings, 
and therefore the work should be intrusted only to those who have 
given careful study to the matter and who have had extended experience 
in burning the crude product. The form of the furnace will play a very 
small part in increasing the use of crude petroleum. The method and 
character of the installation will count for much, but where burners are 
simple in design and are constructed in accordance with scientific 
principles there will be very little difference in their efficiency. Con- 
sumers should principally see that they do not purchase appliances that 
have been untried and have been designed by persons who have had 
but limited experience in operating oil devices. 

57. Gaseous Fuels. — These fuels offer all of the advantages of liquid 
fuels and but few of the disadvantages. The gases most commonly met 
with in connection with steam power plants are outlined in Table 11. 
The artificial gases for steam purposes are prohibitive in cost in 
most cases, and even in blast-furnace installations, where the gases are 
waste products, the gas engine has virtually supplanted the steam engine 
for power purposes. In the immediate locality of natural gas wells gas- 
fired furnaces may prove to be more economical than coal furnaces, but 
the limited supply and constant fluctuation in price limit its use as a 
general fuel. From the market quotations on coal and gas and the 
comparative heating value of each the actual cost per B.T.U. is readily 
obtained, and by combining this with the relative efficiencies from the 
furnace standpoint the net cost of the fuel is obtained. The following 
table, based upon the assumption that one cubic foot of natural gas 
under standard conditions has a heating value of 1,000 B.T.U. , will 
enable an approximate comparison to be made: 






FUELS AND COMBUSTION 



67 



B.T.U. per Pound of 
Coal. 


Pounds of Coal Equal 

to 1,000 Cu. Ft. 

of Gas. 


No. of 1,000 Cu. Ft. 

of Gas Equal to One 

Short Ton of Coal. 


10,000 
11,000 
12,000 
13,000 
14,000 
15,000 


100 

91 
83 
77 
71 
67 


20 
22 
24 
26 

28 
30 



Fuel Economy : Fuel Economy in Steam Power Plants : Cassier's Mag., May, 1904; 
Inst, of Elec. Engrs., Jan. 12, 1905; Engr. U.S., April 1, 1905; Engng., Aug. 7, 1903; 
Eng. Mag., June, 1907. The Province of the Fuel Expert : Eng. and Min. Jour., May 25, 
1905. A Gas-Fired Boiler : Engr. U.S., Feb. 15, 1907, p. 223. 

See also A.S.M.E. Code for conducting Boiler Tests — reprinted in Appendix B. 



TABLE 11. 

CHARACTERISTICS OF GASEOUS FUELS. 

(Lucke.) 



Natural gas 

Cannel-coal gas 

Common-coal gas 

Carburetted water-gas. . . 
Uncarburetted water-gas. , 
Producer-gas, little steam. 
Loomis Pettibone coal gas 

Dowson gas, average 

Taylor gas, average 

Mond gas , 

Coke-oven gas 

Blast-furnace gas , 



H 



1.7 
27.7 
39.78 
21.8 
49.50 

9.2 
14.0 
18.0 
21.0 
29.0 
53.0 

3.0 



CO 



0.55 

6.8 

7.04- 
28.1 
35.93 
25.3 
20.0 
25.0 
12.0 
12.0 

6.0 
27.5 



CH 4 



94.16 

50.0 

45.16 

30.7 

1.05 

3.1 

2.0 

3.0 

2.0 

2.0 

35.0 



C 2 H 4 



0.30 
13.0 

6.38 
12.9 



0.8 
0.20 



2.0 



0.30 



0.06 
0.5 



0.10 
3.6 





C0 2 


N 


Cubic Feet 

of Air per 

Cubic Foot 

of Gas. 


B.T.U. per Cubic 

Foot 

of Gas. 




High. 


Low. 


Natural gas 


0.29 

0.1 

1.08 

3.8 

4.25 

3.4 

8.2 

7.0 

6.0 

14.5 
2.0 

10.0 


2.80 
2.4 
0.50 
2.2 
8.75 
58.2 
55.5 
47.0 
57.0 
42.5 
2.0 
59.4 


9.13 
6.50 
6.38 
6.00 
2.10 
1.24 


989 

843 
727 
702 
295 
160 


888 


Cannel-coal gas 


762 


Common-coal gas 


651 


Carburetted water-gas 

Uncarburetted water-gas 

Producer-gas, little steam 

Loomis Pettibone coal gas .... 
Dowson gas, average 


635 
265 
150 


1.32 
0.98 
1.17 
5.06 
.81 


119 
130 
156 
620 
100 


115 


Taylor gas, average 


116 


Mond gas 


139 


Coke-oven gas 


524 


Blast-furnace gas 


99 



CHAPTER III. 

BOILERS. 

58. As affecting fuel economy the boiler equipment is by far the 
most important part of the power plant and involves the largest 
share of the operating expenses. It matters little how elaborate, 
modern, or well designed it may be, skill, good judgment, and con- 
tinued vigilance are required on the part of the operator to secure the 
best efficiency. 

Of the various types and grades of boilers on the market experience 
shows that most of them are capable of practically the same evaporation 
per pound of coal, provided they are designed with the same proportions 
of heating and grate surface and are operated under similar conditions. 
They differ, however, with respect to space occupied, weight, capacity, 
first cost, and adaptability to particular conditions of operation and 
location. 

59. Classification. — As to design and construction there is an almost 
endless variety of boilers and furnaces, classified as internally and 
externally fired; water tube and fire tube; through tube and return tubular; 
horizontal and vertical. 

The internally fired type includes the vertical tubular, locomotive, 
Scotch-marine, and practically all flue boilers. The externally fired 
includes the plain cylinder, the through tubular, return tubular, and 
nearly all stationary water-tube boilers. 

60. Vertical Tubular Boilers. — Vertical tubular boilers, Figs. 1 and 
23, are commonly used where small power, compactness, low first cost, 
and sometimes portability are the chief requirements, though they are 
not necessarily restricted to small sizes. The tubes are sometimes 
arranged so that the spaces between them radiate from a hand hole on 
one side so that a scraper may readily be inserted to clean the top of the 
furnace plate. The hand hole in the water leg permits removal of the 
scale. It is convenient to place a chain in the bottom of the water leg 
which can be worked around through the hand hole for the purpose 
of loosening up the scale deposit. The distance between the furnace 
crown and top of the grate is never less than 24 inches even in the smallest 
boiler and should be as great as possible to insure good combustion. 

/Two styles of vertical boilers are in common use, the ordinary vertical 

68 



BOILERS 



69 



type, Fig. 1, and the submerged type, Fig. 23. In the former the upper 
tube sheet and part of the tubes are above the water line, and while this 
feature may tend to superheat the steam to a slight extent, the difficulty 



STACK 



STEAM GAUGE 



HAND-HOLE 




WATER COLUMN 



STAY-BOLTS 



BLOW OFF 



Fig. 23. Vertical Tubular Boiler with Submerged Tube Sheet. 



from unequal expansion and liability to overheating is of sufficient 
moment to justify the use of the submerged type, particularly where the 
boiler is likely to be forced above its rated capacity. The advantages 
of this type of boiler are (1) compactness and portability; (2) requires 
no setting beyond a light foundation; (3) is a rapid steamer, and (4) is 
low in first cost. The disadvantages are (1) inaccessibility for thorough 
inspection and cleaning; (2) small steam space, which results in excessive 



70 



STEAM POWER PLANT ENGINEERING 



priming at heavy loads; (3) poor economy except at light loads, as the 
products of combustion escape at a high temperature on account of the 





SECTION THROUGH A- 




SECTION THROUGH ASH PIT 



SECTIONAL FRONT ELEVATION 



Fig. 24. Manning Vertical 
Fire-Tube Boiler. 



shortness of the tubes; (4) smokeless combustion practically impossible 
with bituminous coals; (5) the small water capacity results in rapidly 
fluctuating steam pressures with varying demands for steam. 






BOILERS 



71 



Although vertical fire-tube boilers are usually of very small size, 
being seldom constructed in sizes over 60 horse power, an exception is 
found in the Manning boiler, Fig. 24, which is constructed in sizes as 
large as 250 horse power. Many of the disadvantages found in the 
smaller types are obviated in the Manning boilers, which, as far as 
safety and efficiency are concerned, rank with any of the other first- 
class types. They differ from the boiler described above mainly in 
having the lower or furnace portion of much greater diameter than the 
upper part which encircles the tubes. This permits a proper proportion 
of grate, which is not obtainable in boilers like Figs. 1 and 23. The 
double flanged head connecting the upper and lower shells allows 
sufficient flexibility between the top and bottom tube sheets to provide 
for unequal expansion of tubes and shell. The ash pit is built of brick 
and the water leg does not extend below the grate level, thus doing away 
with dead water space. Where overhead room permits and ground 
space is expensive, this boiler offers the advantage of taking up a small 
floor space as compared with horizontal types. 

61. Fire-Box Boilers. — Although vertical fire-tube boilers may be 
classed as fire-box boilers, yet the term "fire box" is usually associated 
with the locomotive types, whether used for traction or stationary pur- 
poses. The usual form of fire-box boiler as applied to stationary work 



SAFETY VALVE 




FIRE DOOR 



ASH DOOR 



Fig. 25. Typical Fire-box Boiler. — Stationary Type. 



is illustrated in Fig. 25. The shell is prolonged beyond the front tube 
sheet to form a smoke box. The front ends of the tubes lead into the 
smoke box and the rear ends into the furnace or fire box. The fire box 



72 



STEAM POWER PLANT ENGINEERING 



is ordinarily of rectangular cross section, and is secured against collapse 
by stay bolts and other forms of stays. In Fig. 25 the smoke box is of 
cylindrical cross section and hence requires no staying except at the 
flat surface. Fire-box boilers are used a great deal in small heating 
plants where space limitation precludes other types. Their steam 
capacity gives them an advantage over the vertical tubular form. 
Being internally fired no brick setting is required. They are usually 
of cheap construction, designed for low pressure, and seldom made in 
sizes over 75 horse power. Unless carefully designed and constructed 
high steam pressures are apt to cause leakage because of unequal expan- 
sion of boiler shell, tubes, and fire box. Portable fire-box boilers with 
return tubes are made in sizes as large as 150 horse power and for 
pressures as high as 150 pounds per square inch, but being more costly 
than some of the other types of boilers of equal capacity are used only 
where portability is an essential requirement. 

62. Scotch-Marine Boiler. — Where an internally fired boiler is 
desired for large powers the Scotch-marine type is finding much favor 
with engineers. A number of the tall office buildings in Chicago are 
equipped with boilers of this class which are giving good results. They 
require little overhead room, no brick setting, and are excellent steamers. 




mt»»w»>w»jwj»»»»»w»>»iMw»w»wm/M»»MM»m 



r ~ 



Fig. 26. Stationary Scotch-Marine Boiler. 



The Continental boiler, Fig. 26, is one of the best known of this type. 
The boiler is self-contained and requires no brick setting, the only fire 
brick used being those that form the bridge wall, baffle ring and the 
layer at the back of the combustion chamber. The furnace and tubes are 



BOILERS 



73 



entirely surrounded by water, so that all fire surfaces, excepting the 
rear of the combustion chamber, are water cooled. The furnace is cor- 
rugated for its whole length. These corrugations, in addition to giving 
greater strength to the furnace, act as a series of expansion joints, 
taking up the strains due to unequal expansion of furnace and shell. 
Practically all types of mechanical stokers and grates are applicable to 
these boilers. The advantages of a Scotch boiler and of all internally 
fired boilers are (1) minimum radiation losses; (2) requires no setting; 
(3) no leakage of cool air into the furnace as sometimes occurs through 
cracks or parous brickwork of other types; (4) large steaming capacity 
for the space occupied. The circulation, however, is not always positive 
and the water below the furnace may be considerably below the average 
or normal temperature, giving rise to unequal expansion and contraction 
which may cause leakage. The boiler proper is relatively costly, but 
this is offset to some extent by the absence of setting. 

63, Robb-Mumford Boiler. — Fig. 27 shows a section through a 
Robb-Mumford boiler, which is a modification of the Scotch-marine 




and of the horizontal tubular type. It consists of two cylindrical 
shells, the lower one containing a round furnace and tubes and the 
upper one forming the steam drum, the two being connected by two 
necks. The lower shell has an incline of about one inch per foot from 
the horizontal, for the purpose of promoting circulation and draft, 
and also for convenience in washing out the lower shell. Combustion 
takes place in the furnace, which is surrounded entirely by water, and 



74 STEAM POWER PLANT ENGINEERING 

the gases pass through the tubes and return between the lower and 
upper shells (this space being enclosed by a steel casing) to the outlet 
at the front of the boiler. Mingled water and steam circulate rapidly 
up the rear neck into the steam drum, where the steam is released, the 
water passing along the upper drum towards the front of the boiler and 
down the front neck, a semicircular baffle plate around the furnace 
causing the down-flowing water to circulate to the lowest part of the 
lower shell under the furnace. The outer casing, which incloses the space 
between the lower and upper shells, including the rear smoke box and 
the smoke outlet, is constructed of steel plate, with angle-irqn stiffeners, 
the various sections being bolted together for convenient removal. The 
inside of the steel case, including the rear smoke chamber, is lined with 
asbestos air-cell blocks fitted in between the angle-iron stiffeners. The 
top of the upper drum and bottom of the lower shell are also covered 
with non-conducting material after the boiler is erected. Owing to 
the fact that steam and water spaces are divided between two cylin- 
drical shells, the thickness of plates is not so great as in the Scotch- 
marine or horizontal return tubular types; and the rear chamber of the 
marine boiler is avoided. 

The chief claim for this type of boiler is compactness. A battery of 
five 200-horse-power units occupies a floor space of but 33 feet in width 
by 20 feet in depth and 12.5 feet high. Each unit is entirely independent 
and may be isolated for cleaning, inspection, and repairs. 

64. Horizontal Return Tubular Boilers. — These are the most common 
in use and are constructed in sizes up to 200 horse power. They are 
simple and inexpensive and, when properly operated, durable and 
economical. Figs. 28 to 31 show various forms of standard settings, 
and Figs. 75, 76, and 76a different " smokeless " settings. The grate is 
independent of the boiler, and the products of combustion pass beneath 
the shell to the back end, returning through the tubes to the front, 
and into the smoke connection. 

The tubes are 3 to 4 inches in diameter and from 14 to 18 feet long, 
and are expanded into the tube sheets. The portion of the tube sheets 
not supported by the tubes is secured against bulging by suitable stays. 
Access to the interior of the boiler is obtained through manholes. The 
most convenient arrangement for inspection and cleaning is to have 
one manhole located at the top of the shell and one at the bottom of 
the front tube sheet. Return tubular boilers are made either with an 
extended front (Fig. 28) or flush front (Fig. 29). The latter costs a 
little more for brick and setting, but it is more convenient to operate 
and the boiler is less expensive. The shell may be supported by lugs on 
the brickwork as in Fig. 28 or by steel beams and hangers as in Fig. 30. 



BOILERS 



75 







9 -.01 




1 

! 




76 



STEAM POWER PLANT ENGINEERING 







BOILERS 



77 



The latter construction permits the brickwork and shell to expand 
or contract independently, and settling of the brickwork does not affect 
the boiler alignment. With the side bracket support, the front lugs 
usually rest directly on iron or steel plates imbedded in the brickwork, 
and the back lugs on rollers, to permit free expansion and contraction. 
The brackets are long enough to rest upon the outside wall, so that the 
inside brick lining can be renewed without disturbing the setting. The 
distance between the rear tube sheet and wall should be about 16 inches 



STEEL SUPPORTS 




Fig. 30. Return Tubular Boiler Setting. — Steel Beam Suspension. 



for boilers less than 60 inches in diameter and from 20 to 24 inches for 
larger ones. The distance between grate and boiler shell should not 
be less than 28 inches for anthracite coal and 36 inches for bituminous 
coal.* The greater this distance the more complete the combustion, 
since the gases will have a better opportunity for combining with the air 
before coming in contact with the comparatively cool surfaces of the 
shell. The shell should be slightly inclined toward the blow-off end 
so as to drain freely. 

The vertical distance between the bridge wall and shell is usually 
between 10 and 12 inches. The lower part of the combustion chamber 
behind the bridge wall may be filled with earth and paved with common 

* For smokeless combustion the setting must be modified. See furnace illustrated and 
described in paragraph 96. 



78 



STEAM POWER PLANT ENGINEERING 




a 

o 
O 



BOILERS 



79 



brick, as in Fig. 31, or left empty as in Fig. 29. The shape of the walls, 
whether curved to conform to the shell or flat, appears to have little 
influence on the economy. 

The side and end walls are ordinarily constructed of common brick 
with an inner lining of fire brick, and may be solid as in Fig. 29 or 
double with air spaces as in Fig. 28. The latter construction is prefer- 
able and permits the inner and outer walls to expand independently 
without cracking and settling. The side walls are braced by five pairs 
of buck-staves, with through rods under the paving and over the tops 
of the boilers. 

The connection between the rear wall and the shell is a source of 
more or less trouble on account of the expansion and contraction of the 
boiler. Cast-iron supports of T section supporting a fire-brick arch 
are usually employed as illustrated in Fig. 32, the clearance between 
the arch and the shell being sufficient to allow the necessary expansion. 




Fig. 32. Furnace Arch Bars. 



Fig. 33. Back connection made with 
Cast-iron Plate. 



Fig. 33 shows the common method of resting one end of the arch 
supports on the rear wall and the other end on an angle iron riveted to 
the boiler. 

The products of combustion are sometimes carried over the top of the 
boiler as shown in Fig. 31. This tends to superheat the steam, but the 
advantage gained is probably offset considerably by the extra cost of 
the setting and the accumulation of soot on the top of the shell. The 
arrangement is not common. 

The steam connection is naturally made to the highest point in the 
boiler shell. Frequently a steam dome, to which the steam nozzle is 
connected, is provided as in Fig. 29. The function of the steam dome is 
to increase the steam space so as to permit the collection of dry steam at 
a point high above the water level. If a boiler is too small for its work 



80 



STEAM POWER PLANT ENGINEERING 




7e « 



BOILERS 



81 



and is forced far above its rating a steam dome is probably an advantage, 
though its use is less common now than formerly, since a properly 
designed boiler insures ample steam space without one. A dry pipe 
inside the boiler above the water line as in Fig. 26 or 27 is commonly 
used to guard against priming where the nozzle is connected to the shell. 

For low pressures and small powers the return tubular boiler has the 
advantage of affording a large heating surface in a small space and 
large overload capacity. It requires little overhead room and its first 
cost is low. On the other hand the interior is difficult of access for 
purposes of cleaning and inspection. Boilers of this type are seldom 
constructed in sizes above 150 horse power or for pressures over 150 
pounds per square inch, since the cost increases rapidly as the pressure 
rises above this amount. 

65. Babcock & Wilcox Boiler. — Fig. 34 shows a longitudinal section 
through a Babcock & Wilcox boiler, illustrating a typical horizontal 
water-tube type. The tubes, usually 4 inches in diameter and 18 feet 



Fig. 35. Details of 
Header, — Babcock 
and Wilcox Boiler. 



Fig 




Front Section, — Babcock and Wilcox Boiler. 



in length, are arranged in vertical and horizontal rows and are expanded 
into pressed-steel headers. Two vertical rows are fitted to each header 
and are " staggered " as shown in Fig. 35. The headers are connected 



82 STEAM POWER PLANT ENGINEERING 

with the steam drum by short tubes expanded into bored holes. Each 
tube is accessible for cleaning through openings closed by covers with 
ground joints held in place by wrought-iron clamps and bolts. The 
tubes are inclined at an angle of about 22 degrees with the horizontal. 
The rear headers are connected at the bottom to a cast-iron mud drum. 
The steam drum is horizontal and the headers are arranged either ver- 
tically or at right angles to the tubes. The boiler is supported by steel 
girders resting on suitable columns independent of the brick setting. 
The grate is placed under the higher ends of the tubes, the products of 
combustion passing at right angles to the tubes and being deflected 
back and forth by fire-tile baffles. The feed water enters the front of 
the steam drum as shown in Fig. 36. A rapid circulation is effected by 
the difference in density between the solid column of water in the rear 
header and the mixed steam and water in the front one. B. & W. 
boilers under 150 horse power have but one steam drum, and the larger 
sizes have two. The number of tubes varies with the size of boiler, 
ranging from 6 in width and 9 in height in the 100-horse-power boilers 
to 14 high and 18 wide in the 500-horse-power boilers. 

66. Heine Boiler. — Fig. 37 shows a longitudinal section through a 
Heine horizontal water-tube boiler. This boiler differs from the B. & W. 
boiler in that the tubes are expanded into a single large header con- 
structed of boiler steel. The drum and tubes are parallel with each 
other and inclined about 22 degrees with the horizontal. The feed 
water enters at the front of the steam drum and flows into the mud 
drum, from which it passes to the rear header. Steam is taken from 
the front of the steam drum and is partially freed from moisture by the 
dry pipe A. A baffle over the front header prevents an excess of water 
from being carried into the dry pipe. As the rear header forms one large 
chamber, no additional mud drum is necessary and the sediment is 
blown off from the bottom by the blow-off cock. The circulation is 
somewhat freer than in the B. & W. boiler on account of the large 
sectional area through the headers. 

67. Wickes Boiler. — Fig. 38 shows a section through a Wickes 
vertical boiler, illustrating the vertical water-tube type. The steam 
drum and water drum are arranged one directly above the other. The 
tubes are expanded and rolled into both tube sheets and are divided 
into two sections by fire-brick tile. The water line in the steam drum 
is carried about two feet above the tube sheet, leaving a space of five 
feet between water line and top of the drum. This affords a large 
steam space and disengagement surface. Feed water is introduced 
into the steam drum below the water line and flows downward through 
the tubes of the second compartment. The boiler is supported by four 



BOILERS 




84 



STEAM POWER PLANT ENGINEERING 



brackets riveted to the shell of the bottom drum and is independent 
of the setting. The entire boiler is enclosed in brickwork and is com- 
pletely surrounded by the products of combustion. The upper part 





llil 



1 i i 





Fig. 38. Wickes Vertical Water Tube Boiler. 



of the steam drum acts as a superheating surface and tends to dry 
the steam. Wickes boilers are simple in design, easy to inspect and 
clean, low in first cost, and comparable in efficiency with any water- 
tube type of boiler. 



BOILERS 



85 




86 



STEAM POWER PLANT ENGINEERING 



a 



5.9 



67a. Parker Boiler. — Fig. 38a shows a longitudinal sectional ele- 
vation and an end sectional elevation of a 1200-H.P. Parker Down- 
Flow Boiler with double-ended setting. This type of boiler is finding 
much favor with engineers for central stations where large units are 
desired. The Parker boiler differs from the conventional horizontal 
water-tube boiler principally in circulation and flexibility. 

Feed water is pumped into the economizer or feed element (1), Fig. 
38a, at 0, 0, and flows downward through a series of tubes, discharging 
finally into the drum through an upcast H. In a large unit, as illus- 
trated here, there are two feed elements and two drums. The circula- 
tion in the feed element 
is indicated by solid 
lines and arrow points 
at the left of the end 
sectional elevation, the 
tubes having been omit- 
ted from the drawing 
for the sake of clearness. 
The intermediate ele- 
ments (2) take their 
water supply from the 
bottom of the drum 
through a cross-box V, 
the circulation being 
downward, as indicated 
by arrow points, 
through four tube wide elements, and finally discharge it through an 
upcast X into the steam space of the drum. Each element has a 
" down-comer " and an upcast. In the smaller sized boilers the inter- 
mediate elements are omitted. 

The evaporator elements (3) take their water supply from the bottom 
of the drum at V, the circulation being downwards through two tube 
wide elements, and finally discharge it into the drum at U. The last 
two passes of the water are through the two bottom tubes of each ele- 
ment, thus assuring dry steam without the use of dry pipes. To prevent 
reversal of flow each element is fitted with a check valve at the admission 
end. Each drum is equipped with a diaphragm, as indicated, separating 
the steam and water spaces, thus insuring against foaming and priming. 
Saturated steam is taken from the drum at A and passes by way of 
B to C, where it enters the superheater S. The superheated steam 
leaves the superheater at D and passes by way of E and R to the storage 
drum N, finally leaving the boiler at G. The superheater is designed 

















1 










1 


1 


BOILER R 
SIZE? OF 


OOM AREA FOR VARIOUS TYPES AND 
BOILERS, 2 BOILERS IN A BATTERY 1 




















1 
1 


























B 4 W- 3,5,7,8,9,10,11,12,13,11,15,16 

Sterling-2,4,6 

Helne-1 

Parker-20,21 

Hormaby-18,19 




























































1 




5 < 


o 11 
5 4 


























I x 









15 


























6^ 


3 R 




12 
































13 








in 






























t 











































































500 



1000 1500 

Capacity of Battery Boiler H.P. 

Fig. 38b. 



2000 



BOILERS 



87 



to maintain an approximately constant degree of superheat for all 
variations in load. 

All tubes are connected by malleable-iron junction boxes, the interior 
of each tube being accessible through hand holes placed opposite the 




Fig. 39. Stirling Boiler and Setting. 



end of each tube. The hand-hole cover plates are on the inside of the 
box and have conical ground joints, thus dispensing with gaskets. 

The Parker boiler is built single or double ended, with or without 
superheater, and in sizes ranging from 50 H.P to 2500 H.P. standard 
rating. 

68. Stirling Boiler. — Fig. 39 shows a longitudinal section through 
a Stirling water-tube boiler, which differs considerably from the types 
just described. Three horizontal steam drums and one horizontal mud 
drum are connected by a series of inclined tubes. The tubes are bent 
at the ends to permit them to enter the drums radially. Short tubes 



88 STEAM POWER PLANT ENGINEERING 

connect the steam spaces of all the upper drums and also the water 
spaces of the front and middle drums. Suitably disposed fire-tile 
baffles between the banks of tubes direct the gases in their proper 
course. The boiler is supported on a structural steel framework in- 
dependent of the setting. The feed water enters the rear upper drum, 
which is the cooler part of the boiler, and flows to the bottom or mud 
drum, where it is heated to such an extent that many of the impurities 
are precipitated. There is a rapid circulation up the front bank of 
tubes to the front drum, across to the middle drum, and thence down 
the middle bank of tubes to the mud drum. The interior of the drums 
is accessible for cleaning by manholes located in the ends. The Stirling 
furnace is distinctive in design. A fire-brick arch is sprung over the 
grates immediately in front of the first bank of tubes. The large tri- 
angular space between boiler front, tubes, and mud drum forms the 
combustion chamber. Stirling boilers are somewhat lower in first 
cost than other types of water-tube boilers on account of the absence 
of numerous hand holes and the like which are necessary in the hori- 
zontal type. 

69. Unit of Evaporation. — The performance of a boiler and furnace 
may be expressed in terms of the weight of water evaporated per hour 
per square foot of heating surface or of the weight evaporated per pound 
of fuel. To reduce all performances to an equal basis so as to facilitate 
comparison the evaporation under actual conditions is conveniently 
referred to the equivalent evaporation from a feed-water temperature 
of 212 degrees F. to steam at atmospheric pressure. The heat required 
to evaporate one pound of feed water at a temperature of 212 degrees F. 
into steam of the same temperature, or " from and at 212 degrees " 
as it is commonly called, is 965.7 B.T.U.* The ratio of the heat neces- 
sary to evaporate one pound of water under actual conditions of feed 
temperature and steam pressure to the heat required to evaporate one 
pound from and at 212 degrees is called the factor of evaporation. Thus 
for dry steam, 

X _ t + 32 

F = 965.7 ' (U) 

in which 

F = factor of evaporation. 

X = total heat of one pound of steam at observed pressure. 
t = temperature of the feed water, degrees F. 

* Recent redeterminations of the properties of saturated steam give this figure 
as 970.4. 



BOILERS 



89 



Dry Surfaces 



Heatfrom Fuel j 

Bed and Hot > 

Furnace Walls \ 



If the steam is wet, 

X - xr + q, (12) 

in which 

q = heat in liquid at observed pressure. 

x = the quality of the steam. 

r = latent heat of evaporation at observed pressure. 

If the steam is superheated, 

X = r + q+ Ct s , (13) 

in which 

C = the specific heat of the superheated steam. 

t s = the degree of superheat, degrees F. 

69a. Heat Transmission. — Fig. 39a shows a section through a boiler 
heating plate and serves to illustrate the accepted theory of heat trans- 
mission. The outer surface of 
the plate is covered with a thin 
layer of soot and a film of gas, 
and the inner surface is similarly 
protected by a layer of scale and 
a film of steam and water. It is 
therefore reasonable to assume 
that the dry surface of the plate 
is located somewhere within the 
film of gas, and the wet surface 
within the film of water and 
steam. 

The heat is imparted to the dry 
surface by (1) radiation from the 
hot fuel bed and furnace walls, 
and by (2) convection from the 
moving furnace gases. The heat 
is transferred through the boiler 
plate and its coatings purely by 
conduction. The final transfer 
from the wet surface to the boiler 
is mainly by convection. 

Radiation depends on the temperature, and according to the law of 
Stephen and Boltzmann is approximately proportional to the difference 
between the fourth power of the absolute temperature of the fuel bed 
and furnace walls and the temperature of the dry surface of the heating 
plate. According to this law the heat transmitted by radiation increases 
rapidly with the increase in furnace temperature. In the modern boiler 
the surface exposed to radiation is only a small portion of the total 




o- 1 



Fig. 39a. 



A =• Average Temperature of.Moving Gases. 
B= Average Temperature of Dry Surface. 
C —Average Temperature of Wet Surface . 
D ^Temperature of Water in Boiler. 

Heat Transmission through 
Boiler Plate. 



90 STEAM POWER PLANT ENGINEERING 

heating surface, and, since in well-operated furnaces the temperature of 
the furnace cannot be increased materially on account of practical con- 
siderations, there is little hope of increasing the capacity of a boiler by 
increasing the furnace temperature. 

The heat imparted to a boiler plate by convection may be determined 
by the following equation (Prof. Perry, " The Steam Engine," 1906 
Ed., p. 588): 

H = C & - t 2 ) vol, (13a) 

in which ^ 

H = B.T.U. transferred per hour per sq. ft. of heating surface. 

C = a coefficient determined by experiment. 

t x = temperature of the moving gases, degrees F. 

t 2 = temperature of the dry plate surface, degrees F. 

v = velocity of the gases, feet per sec. 

d = density of the gases, lbs. per cubic foot. 

Prof. Nicholson gives the following modifications of formula (13a) 
as applied to boiler tubes or flues (Engr. Lond., Feb. 19, 1908): 



in which 



H -\m + wt{ l + i)V k - Qvd ' (13b) 



t = mean film temperature. 

m = hydraulic radius = area of tube in square inches -*- perimeter 
of the tube in inches, other notations as in (13a). 

Both equations are based upon the same general law except that the 
latter gives a means of determining coefficient C in terms of the mean 
film temperature and the dimensions of the flues or tubes. 

An examination of equation (13a) shows that for a given set of condi- 
tions the heat imparted to a unit of dry surface of heating plate varies 
directly as the difference between the temperature of the hot gases and 
that of the dry surface and directly as the velocity and density of the 
gases. However, the density of the gases drops with the rise of tem- 
perature, and increase in furnace temperature does not necessarily imply 
increase in heat impartation. It is the utilization of the velocity factor, 
then, which offers a possibility of increasing boiler capacity and efficiency. 

Experiments by Prof. Nicholson and the U. S. Geological Survey show 
that by establishing a powerful scrubbing action between the gases and 
the boiler plate the protecting film of gas is torn off as rapidly as it is 
formed and new portions of the hot gases are brought into contact with 
the plate, thereby greatly increasing the rate of heat transmission. 
Similarly the faster the circulation of the water the greater will be the 



BOILERS 



91 



scrubbing action tending to remove the bubbles of steam from the wet 
surface and the more rapid will be the transfer from the plate to the 
boiler water. 

The resistance of the metal itself is so small that it may be neglected 
in calculating the heat trans- 
mission, and it may be logically 
assumed that the plate will take 
care of all the heat that reaches 
its dry surface. 

Prof. Nicholson found that by 
filling up the flue of a Cornish 
boiler with an internal water 
vessel, leaving an annular space 
of only 1 inch around the latter, 
an evaporation eight times the 
ordinary rate was effected at a 
flow of gases 330 feet per second 
(8 to 10 times the average flow). 
The fan for creating the draft 
consumed about \\% of the 
total power. 

The conclusion is that the 
heating surface for a given evap- 
oration at the present rating 
may be reduced as much as 90% 
for the same output, with a cor- 
responding reduction in the size, 
cost and space requirements, or 
with a given heating surface of 
standard rating the output may 
be enormously increased; also 
the increase in power necessary 
to create the draft is by no 
means comparable with the ad- 
vantages gained. 

The modern locomotive boiler 
is the nearest approach to these 
conditions in practice. Here a powerful draft forces the heated gases 
through small tubes at a very high velocity and an enormous evapo- 
ration is effected with a comparatively small heating surface. See 
Fig. 39b for influence of draft on the capacity of a torpedo boat boiler 
(Power and Engr., May 24, 1910). 



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100 200 300 

Per cent, of Rated Capacity Developed by Boiler. 

Power 

Fig. 39b. Influence of Draft on the Capacity 
of a Normand Water-Tube Boiler on the 
U. S. Torpedo Boat " Biddle." 



92 STEAM POWER PLANT ENGINEERING 

These principles have been applied to a limited extent to stationary 
boilers already installed by making the gas passages smaller as compared 
to the length by means of suitable baffles (Fig. 38a) and by forcing 
larger weights of gas through the boiler, either by forced draft or by 
increasing the grate area (Fig. 68a). 

In a general sense when the capacity of a boiler is doubled or tripled 
the over-all efficiency of the whole steam-generating apparatus drops, 
but the advantage gained usually offsets the loss in fuel economy. A 
close examination of the results, however, will show that the loss in 
efficiency is due more to low furnace efficiency than to inability of the 
boiler to absorb the heat generated. 

In view of recent experiments it is not unlikely that within the next 
ten years boilers will be constructed capable of developing a boiler horse 
power with two or three square feet of heating surface instead of ten 
square feet, as at present, and with high over-all efficiency. (See 
Figs. 41a and 41b.) 

Heat Transmission in Boilers, Kreisinger and Ray: Power and Engr., June 29, 
1909, p. 1144; U. S. Geological Survey, Bulletin Journ. West Soc. Engrs., Sept. 
18, 1907; Am. Inst. Elect. Engrs., Dec. 13, 1907. 

Heat Transfer and Future Boiler Practice : A. H. Allen, Power and Engr., Sept. 
21, 1909, p. 482; Engng., Lond, Feb. 19, 1908. 

The Heat of Fuels and Furnace Efficiency : W. D. Ennis, Power and Engr., r July 
14, 1908, p. 50. 

A Study in Heat Transmission (The Transmission of Heat to Water in Tubes as 
Affected by the Velocity of the Water), J. K. Clement and C. M. Garland, Univ. 
of 111. Bulletin No. 40, Sept. 27, 1909. 

70. Heating Surface. — All parts of the boiler shell, flues, or tubes 
which are covered by water and exposed to hot gases constitute the 
heating surface. Any surface having steam on one side and exposed 
to hot gases on the other is superheating surface. According to the 
recommendations of the American Society of Mechanical Engineers, 
the side next to the gases is to be used in measuring the extent of the 
heating surface. Thus measurements are made of the inside area of fire 
tubes and the outside area of water tubes. The heating surface in a 
boiler under average conditions of good practice is most efficient when 
the heated gases leave the uptake at a temperature of 100 to 200 
degrees F. above that of the steam. Each square foot of heating surface 
is capable of transmitting a certain amount of heat, depending upon the 
conductivity of the material, the character of the surface, the temperature 
difference between the gas and the water, the location and arrangement of 
the tubes, the density of the gas, the velocity of the gas, and the time allowed 
for transmission of the heat. It is customary to assume a uniform heat 



BOILERS 93 

transmission for the entire surface. Thus with most boilers it is found 
that the best results are obtained with an evaporation of from 3 to 3.5 
pounds of water from and at 212 degrees F. per square foot of heating 
surface, which is equivalent to allowing 12 to 10 square feet per boiler 
horse power. By increasing the quantity of heat the evaporation may 
be increased, but at the expense of efficiency, since a smaller percentage 
of the heat is utilized. For example, an evaporation as high as 20 
pounds per square foot per hour has been effected in torpedo-boat prac- 
tice, and 12 pounds per square foot per hour is not unusual in locomotive 
work, but such performances are invariably obtained at the expense of 
economy. The selection of the proper proportion of heating surface to 
the evaporation required is evidently a very important matter. For 
maximum economy under average conditions of operation, practice 
allows a proportion of 1 square foot to every 3.5 pounds of water to be 
evaporated from and at 212 degrees F. Where economy must be sacri- 
ficed to capacity, as in locomotive practice, a much higher evaporation 
is allowed. 

The maximum evaporation is limited by the amount of coal which can 
be burned upon the grate. It the draft is sufficient, a good boiler can 
develop a horse power upon 0.75 to 0.5 of the surface recommended. 
In the very latest large central stations the gas passages and grate 
surface are proportioned so that the boiler may be operated at 100% 
above standard rating with high over-all efficiency. 

The following table shows approximately the result which may 
be expected with different rates of evaporation. 



POUNDS WATER EVAPORATED FROM AND AT 212 DEGREES F. PER SQUARE 
FOOT OF HEATING SURFACE PER HOUR. 



2 


2.5 


3 


3.5 


4 


5 


6 


8 


10 


12 


PROBABLE RELATIVE ECONOMY. 


100 


100 


100 


100 


99 


98 


95 


90 


85 


80 



Efficiency of Boiler Heating Surface: Trans. A.S.M.E., 18-328, 19-571. Kent, Steam 
Boiler Economy (John Wiley & Son), Chapter IX. The Nature of True Boiler Effi- 
ciency: Jour. West. Soc. Engrs., Sept. 18, 1907. Heat Transference through Heating 
Surface: Engineering, 77-1. 



71. The Horse Power of a Boiler. — A boiler horse power is equivalent 
to the evaporation of 34.5 pounds of water per hour from a temperature 
of 212 degrees F. to steam at atmospheric pressure. This corresponds 



94 STEAM POWER PLANT ENGINEERING 

to 33,305 B.T.U. per hour.* Since the power from steam is developed 
in the engine and the boiler itself does no work, the above measure of 
capacity is merely conventional. Thus one boiler horse power will 
furnish sufficient steam to develop about three actual horse power in 
the best compound condensing engine, but only one-half horse power 
in a small non-condensing engine. Boilers should be purchased on the 
basis of heating surface and not on the horse power rating, since one 
bidder may offer a boiler with say 5 square feet of heating surface per 
horse power and another with 10 square feet, both being capable of 
the required evaporation, but the one with a small heating surface 
(which will, of course, be the cheaper boiler) will do so only at an 
increased cost of fuel. Manufacturers ordinarily rate their boilers on 
the basis of 10 to 12 square feet of heating surface per horse power, and 
the power assigned is called the builder's rating. As this practice is not 
uniform, bids and contracts should always specify the amount of heating 
surface to be furnished. According to the recommendations of the 
American Society of Mechanical Engineers, " A boiler rated at any 
stated capacity should develop that capacity when using the best coal 
ordinarily sold in the market where the boiler is located, when fired by 
an ordinary fireman, without forcing the fires, while exhibiting good 
economy. And further, the boiler should develop at least one-third 
more than stated capacity when using the same fuel and operated by 
the same fireman, the full draft being employed and the fires being 
crowded; the available draft at the damper, unless otherwise under- 
stood, being not less than one-half inch water column. 

In determining the boiler horse power required for a given engine 
horse power it is convenient to estimate the steam consumption of 
the engine under actual conditions and then ascertain the equivalent 
evaporation from and at 212 degrees F. For example, assume a single 
non-condensing engine developing 20 horse power to use 50 pounds of 
steam per horse power hour, or 1000 pounds steam per hour; steam pres- 
sure, 80 pounds per square inch; feed-water temperature, 120 degrees F. 
Required the boiler horse power necessary to furnish this quantity of 
steam. 

From equation (11), the factor of evaporation is 

-, X-t + 32 1185.3-120 + 32 . 1Q1 
F = 970.4 = ~9704 = U3L 

One thousand pounds of steam under the given conditions are there- 

* With the new value of r = 970.4 in place of 965.7 this figure becomes 
33,478.8. 



BOILERS 



95 



fore equivalent to 1000 X 1.131 = 1131 pounds from and at 212 

degrees F. 

The boiler horse power necessary to furnish steam for the 20-horse- 

power engine will be 

1131 
Boiler horse power = ^—^- = 32.8. 

Example : A 15,000 kilowatt steam turbine and auxiliaries require 
14.7 pounds of steam per kilowatt-hour at rated load; steam pressure 
200 pounds per square inch gauge; superheat 150 degrees F.; feed-water 
temperature, 179 degrees F. 

Required the boiler horse power necessary to furnish this quantity 
of steam. 

The heat furnished to the turbine and auxiliaries per kilowatt-hour is 

w \ X + Cpt s - (t - 32) } = 14.7 { 1199.2 + 0.57 X 150 - (179 - 32) } 
= 16,724 B.T.U. 

t> -i u 15 > 000 X 16 > 724 T^nn , \ 

Boiler horse power = 7Q = 7500 (approx.). 

Table 12 gives the required hourly evaporation per boiler horse power 
at various feed temperatures and steam pressures. 

The following table shows approximately the relation between boiler 
horse power and heating surface for different ratios of evaporation: 





EVAPORATION FROM AND AT 212 DEGREES F. PER SQUARE FOOT 

PER HOUR. 




2 


2.5 


3.0 


3.5 


4 


5 


6 


7 


8 


9 


10 


SQUARE FEET HEATING SURFACE REQUIRED PER HORSE POWER. 


17.3 13.8 11.5 


9.8 


8.6 


6.8 


5.8 


>.9 


4.3 


3.8 


3.5 



Builders of return tubular and vertical fire-tube boilers allow 11 to 12 
square feet of heating surface per horse power; water-tube boilers are 
rated at 10 square feet per horse power, and Scotch-marine boilers at 
8 square feet per horse power. 

72. Grate Surface. — The amount of fuel which can be burned per hour 
limits the amount of water evaporated per unit of time and depends 



96 



STEAM POWER PLANT ENGINEERING 



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BOILERS 97 

upon the extent and nature of the grate surface, the character of the 
fuel, and the draft. A liberal allowance of grate surface is usually 
desirable, particularly when the boilers are to be forced, since too small 
a grate increases the labor of handling and cleaning fires and results 
in poor economy. 

With good coal low in ash approximately equal results may be 
obtained with large grate surface and light draft and with small grate 
and strong draft, the amount of coal burned per hour being the same 
in both cases. Bituminous coal low in ash gives best results with 
high rates of combustion, provided the ratio of grate surface to heating 
surface is properly proportioned. Coals high in ash require a compara- 
tively large grate surface, particularly if the ash is easily fusible, tending 
to choke the grate. Where a strong draft is available a smaller grate 
may be used than with moderate draft, as a thicker bed of fuel can be 
carried. The relation between draft and rate of combustion for various 
sizes of coals is shown in Fig. 116, paragraph 127. 

A number of boiler tests made by Barrus (" Boiler Tests ") showed that 
the best economy with anthracite coal hand fired was obtained with an 
average ratio of grate surface to heating surface of 1 to 36 and at a rate 
of combustion of approximately 12 pounds of coal per square foot of 
grate surface per hour. In these tests a variation in grate and heating 
surface ratio of 1 to 36 up to 1 to 46 gave practically no difference in 
economy. With bituminous coal the tests showed that an average 
ratio of 1 to 45 gave the best results and at a rate of combustion of 
24 pounds of coal per square foot of grate surface per hour. 

Tests made by Christie (Trans. A.S.M.E., 19-330) gave an average 
combustion of 13 pounds of anthracite per square foot of grate per hour 
for maximum efficiency and 24 pounds of bituminous. 

Current central station practice gives normal rates of combustion 
approximately as follows (lbs. per sq. ft. per hr.) : 

Anthracite 15-20 Eastern bituminous 20-24 

Semi-bituminous 18-22 Western bituminous 30-35 

Table 13 gives the relation between heating and grate surface in a 
number of recent boiler installations using different kinds of coal, and is 
illustrative of current practice. 

In proportioning the grate surface for a proposed installation the 
principal factor considered is the character of the fuel, a study being 
made of the various fuels available, and the one selected which gives 
the highest evaporation per dollar. The latter data may usually be 
obtained from records of plants using the same grade of fuel and grates 
similar to those intended for the proposed plant. 



98 



STEAM POWER PLANT ENGINEERING 



TABLE 13. 

RATIO OF HEATING SURFACE TO GRATE SURFACE IN RECENT BOILER 

INSTALLATIONS. 



Nature of Plants. 



Central stations 



Do 

Do 

Do 

Manufacturing 

plants 
Office buildings 



Central station* 



No. of 

Plants. 



10 



9 
20 

6 

1 



Type of 
Boiler. 



Hor water 
tube. 

...do 

...do 

...do 

Return 

tubular. 

..do 



Babcock 
& Wilcox. 



Type of 
Grate. 



Chain 



Roney 

Murphy. . . 
Miscel's . . . 
Hand fired 

Shaking 
grates. 
Roney 



Height of 
Chimney. 



200 feet and 
over. 

...do 

...do...... 

...do 

150-175 

Over 200 . . 
Over 200 . . 



Character of 
Fuel. 



111. screen- 
ings, 15 to 
20% ash. 

Bituminous. 
do. 

Anthracite. 

Anthracite. 

Bituminous 
Bituminous. 



Ratio of 
Heating 
to Grate 
Surface. 



65 



60 
60 
40 
35 

48 

31 



* Two stokers, one at front and one at rear of setting. {Power, Jan. 7, 1908, p. 25.) 

73. Boiler and Furnace Efficiency. — The efficiency of the boiler, 
including the grate, is expressed by the ratio between the heat absorbed 
by the boiler per pound of dry coal fired and the calorific value of one 
pound of dry coal. The efficiency of the boiler alone is taken as the 
ratio between the heat absorbed per pound of combustible burned 
on the grate and the calorific value of one pound of combustible. The 
combustible burned on the grate is equal to the coal as fired minus 
moisture and the total refuse in the ash pit. The calculation of these 
efficiencies is illustrated by the following example: 

ANALYSIS OF COAL. 

Per Cent. 

Moisture 8 

Ash 12 

Combustible 80 

100 

Pounds. 
Water evaporated from and at 212° F. per pound of coal as fired . . . 8.281 

Per Cent. 

Total refuse in ash pit 16 

Percentage of ash in refuse 13 

Combustible in ash 3 

B.T.U. 

Heating value per pound of coal as fired 11,680 

Heating value per pound of dry coal = 11,680 -*- 0.92 12,696 

Heating value per pound of combustible = 11,680 -r- 0.80 14,600 

8.281 + 0.92 = 9.001 = equivalent evaporation per pound of dry coal. 
9.001 X 965.7 = 8,692 = heat absorbed per pound of dry coal. 



BOILERS 99 

Efficiency of boiler and grate = .. ' ana = 68.49 per cent. 

Combustible burned on grate = 100 — (8 + 16) = 76 per cent. 
8.281 -*- 0.76 = 10.896 = equivalent evaporation per pound of com- 
bustible burned on the grate. 
10.896 X 965.7 = 10,522 = heat absorbed per pound of combustible. 

■I f\ coo 

Efficiency of boiler = ' = 72.07 per cent. 

The efficiency of the grate alone might be expressed 

_„. . . Efficiency of boiler and grate 

Efficiency of grate = tW • ti — n — 5 > 

J Efficiency of boiler 

which is equivalent to 

.^~, . , , Combustible actually burned 
Efficiency of grate Combustible fired 

the numerator being the coal fired less moisture and the refuse from 
the ash pit, and the denominator the coal fired less moisture and the 
ash as determined from the proximate analysis. 

The efficiency of combustion is sometimes expressed in terms of the 
difference in temperature between fuel bed and flue gas: 

Tf — T 

Efficiency of furnace = ~ j~ » (14) 

If — la 
in which 

Tf = temperature of the furnace. 
T c = temperature of the flue gas. 
T a = temperature of the air. 
The efficiency of the furnace or combustion may also be stated 
(R. S. Hale, Trans. A.S.M.E., 20-769): 

S + F 
Efficiency of furnace = — = — > (15) 

in which 

S = B.T.U. absorbed by the boiler per pound of dry coal. 

F = B.T.U. lost in flue gases per pound of dry coal. 

H = Calorific value of 1 pound of dry coal. 

The heat absorbed by the boiler expressed in percentage of the heat 
available has been given the name true boiler efficiency by the U. S. 
Geological Survey and may be expressed 

rp rp 

True boiler efficiency = -=£ — ^> (15a) 

J- f J- s 

in which 

T s = temperature of the steam (saturated); other notation as 
in (14). 
74. Boiler Performances. — Table 14 is compiled from a number of 
tests of different types of boilers with various types of grates and 



100 



STEAM POWER PLANT ENGINEERING 



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102 



STEAM POWER PLANT ENGINEERING 



TABLE 14a. 

PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, 

FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. 

(B. & W. Boiler, " Standard " Setting.) 

Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. 
Chain Grate Surface, 90 Sq. Ft. 











H.P. per 
Sq. Ft. 
Grate. 


Heat 


Total 


Super- 


Dry Coal 


Test 


Date, 


Horse 


Eff'y, 


Lost in 


Heating 


heat of 


per Sq. Ft. 


No. 


1908. 


Power. 


PerCent. 


Refuse, 


Surface 


Steam, 


G. S. 










PerCent. 


per H.P. 


Deg. F. 


per Hour. 


2 


Mar. 9 


873 


67.4 


9.70 


2.8 


6.76 


197 


41.2 


4 


" 10 


873 


69.0 


9.52 


2.8 


6.89 


195 


39.1 


6 


" 11 


852 


67.3 


9.47 


2.8 


6.93 


189 


38.9 


8 


" 16 


836 


65.3 


9.29 


6.4 


7.06 


174 


39.5 


10 


a 17 


870 


68.8 


9.67 


5.0 


6.78 


180 


39.3 


14 


" 19 


920 


66.2 


10.22 


9.2 


6.42 


187 


43.7 


16 


" 23 


900 


69.5 


10.00 


4.0 


6.56 


181 


40.5 


18 


" 24 


916 


69.1 


10.18 


5.5 


6.44 


190 


41.6 


20 


" 26 


912 


69.2 


10.13 


4.4 


6.48 


179 


41.2 


22 


" 27 


906 


67.7 


10.07 


4.1 


6.52 


194 


42.5 


24 


" 30 


925 


69.8 


10.28 


2.8 


6.38 


179 


41.6 


26 


" 31 


894 


69.4 


9.93 


5.2 


6.60 


170 


40.6 


28 


Apr. 1 


922 


71.2 


10.24 


3.6 


6.40 


169 


40.4 


30 


" 2 


923 


71.5 


10.26 


4.6 


6.40 


173 


40.5 


32 


u 7 


914 


70.0 


10.20 


4.5 


6.46 


175 


40.9 


34 


" 8 


939 


73.8 


10.4 


3.8 


6.28 


181 


40.4 


36 


" 10 


911 


70.9 


10.1 


3.0 


6.48 


185 


40.2 


38 


" 11 


967 


70.1 


10.7 


3.0 


6.11 


192 


42.6 


40 


" 13 


995 


67.8 


11.1 


3.4 


5.93 


211 


43.6 


42 


" 14 


887 


66.8 


9.9 


4.5 


6.65 


202 


40.8 


44 


" 27 


880 


69.5 


9.8 


5.5 


6.72 


169 


39.7 


48 


" 29 


927 


71.5 


10.3 


3.3 


6.37 


171 


40.8 


50 


" 30 


899 


70.3 


10.0 


4.2 


6.57 


171 


39.6 


52 


May 6 


886 


69.4 


9.8 


5.3 


6.67 


171 


38.2 


54 


" 7 


900 


69.1 


10.0 


4.8 


6.56 


171 


39.2 


56 


" 8 


967 


71.9 


10.7 


4.8 


6.10 


164 


40.1 


58 


" 11 


902 


70.5 


10.0 


3.3 


6.55 


163 


39.6 


60 


" 13 


875 


70.7 


9.7 


3.8 


6.74 


147 


38.3 


64 


" 14 


1102 


72.0 


12.2 


4.8 


5.35 


180 


43.2 



BOILERS 



103 



TABLE 14a. 

PRINCIPAL DATA AND RESULTS OF TESTS ON BOILER NO. 6, UNIT NO. 10, 

FISK ST. STATION. COMMONWEALTH EDISON CO., CHICAGO. 

(B. & W. Boiler, "Standard" Setting.) 

Water-heating Surface, 5000 Sq. Ft. Superheating Surface, 914 Sq. Ft. 
Chain Grate Surface, 90 Sq. Ft. 



Draft 












Heat Lost 






B.T.U. 

per Pound 
Dry Coal. 


Ash in 
Dry Coal, 
Per Cent. 


Ash in 
Refuse, 
Per Cent. 


Uptake 
Temp. 
Deg. F. 


co 2 , 

Per Cent. 


up Stack 
(Dry Gas), 
Per Cent. 


Over 
Fire. 


In 
Uptake. 


.87 


1.34 


11,634 


18.46 


82.33 


466 


6.9 




.78 


1.25 


11,759 


16.81 


81.36 


461 


6.7 




.83 


1.25 


12,039 


16.08 


80.03 


463 


7.7 


15*6 


.94 


1.34 


11,993 


15.91 


67.42 


477 


7.6 


16.8 


.84 


1.24 


11,909 


15.71 


71.32 


475 


7.9 


16.2 


.99 


1.41 


11,768 


16.04 


63.78 


479 


8.5 


15.4 


.77 


1.17 


11,846 


16.68 


79.04 


483 


9.1 


14.0 


.81 


1.25 


11,800 


16.39 


71.98 


484 


8.3 


15.8 


.77 


1.21 


11,846 


15.51 


78.53 


486 


9.0 


14.5 


.78 


1.22 


11,659 


17.59 


80.58 


494 


9.2 


14.6 


.68 


1.28 


11,800 


16.22 


82.97 


487 


8.8 


15.1 


.70 


1.24 


11,752 


16.18 


76.84 


484 


8.8 


15.1 


.62 


1.21 


11,862 


15.38 


82.99 


480 


9.2 


14.1 


.58 


1.40 


11,800 


16.02 


78.37 


480 


9.1 


14.4 


.73 


1.24 


11,815 


16.84 


77.84 


494 


9.0 


14.7 


.72 


1.25 


11,659 


18.06 


82.27 


504 


8.9 


15.3 


.65 


1.13 


11,831 


17.15 


86.92 


493 


9.7 


13.4 


.70 


1.24 


12,002 


16.05 


84.39 


502 


9.0 


15.1 


.71 


1.23 


12,469 


14.87 


82.14 


522 


9.7 


13.3. 


.63 


1.09 


12,049 


15.17 


78.12 


500 


9.5 


13.3 


.71 


1.26 


11,801 


15.75 


77.21 


470 


8.3 


15.7 


.68 


1.23 


11,769 


18.59 


84.04 


472 


8.7 


14.2 


.66 


1.27 


11,955 


16.11 


79.30 


473 


7.9 


16.1 


.62 


1.20 


12,360 


13.63 


74.59 


476 


8.8 


14.5 


.66 


1.31 


12,298 


13.62 


75.19 


480 


9.0 


14.4 


.66 


1.29 


12,423 


13.37 


75.61 


474 


9.4 


13.3 


.92 


1.18 


11,956 


17.45 


83.24 


451 


9.2 


12.5 


.76 


0.98 


11,971 


17.45 


80.99 


443 


10.0 


11.2 


.68 


1.15 


13,126 


10.24 


70.90 


487 


10.4 


12.1 



104 STEAM POWER PLANT ENGINEERING 

characters of fuel. Although some of the tests show a combined 
efficiency of boiler and grate as high as 85 per cent (Engr., Lond., March 
21, 1902, p. 286), such a performance cannot be expected for continuous 
operation under the average conditions of practice. In pumping 
stations or in plants where there are no peak loads and the boiler may 
be operated under a practically constant set of conditions a continuous 
efficiency of 75 per cent has been realized with coal as fuel and 80 per 
cent with crude oil, though these figures are exceptional. In very large 
central stations, with the usual loads in the morning and evening, an 
average efficiency throughout the year of 65 per cent is possible, though 
a good figure is not far from 60 per cent. In large isolated stations with 
variable loads good practice gives an average of 60 per cent. Small 
stations though showing an efficiency as high as 75 per cent at times 
seldom average 50 per cent for the year. The usual discrepancy between 
efficiency as determined by special tests and everyday operation is due 
to the fact that the efficiency test is usually conducted under ideal con- 
ditions: the boiler surfaces are cleaned, the rate of combustion carefully 
adjusted for maximum economy, and special attention given to the firing, 
whereas in actual practice these refinements are seldom attempted. 
Much depends upon the efficiency of the boiler-room staff, the character 
of furnace and fuel, draft, and the load factor. From the commercial 
standpoint the performance is best expressed in terms of the " cost 
to evaporate 1000 pounds of water from and at 212," or the " pounds 
of water evaporated per $1 of coal." Table 15 gives the results of a 
number of tests, made at the Armour Glue Works, Chicago, 111., show- 
ing the cost of evaporating water with different grades of Illinois coal. 
The results were obtained from hand-fired Stirling boilers. 

Boiler Room Economies: Am. Elecn., Oct., 1901, p. 506, Sept., 1905, p. 472; Cas- 
sier's Mag., March, 1906, p. 373; Elec. World, March 4, 1905; Engr. U.S., May 1, 1905, 
p. 304, Jan. 1, 1907; Engr., June 4, 1907, p. 758; Eng. Rec, June 27, p. 685; Elecn., 
Lond., Aug. 5, 1904; Power, Aug., 1905, p. 484; Eng. Mag., Oct., 1901, March, 1903. 

Care and Management of Boilers : Engr. U.S., March 1, 1902, p. 142, Feb. 15, 1904, 
July 15, 1904, Jan. 1, 1907; Engineering, Feb. 18, 1898, p. 211, July 15, 1898, p. 84; 
Eng. Mag., Feb., 1901, p. 877, Oct., 1901, p. 91, March, 1903, p. 896; Power, Sept., 
1904, p. 467, May, 1905, p. 267, Dec, 1905, p. 742, Sept., 1906, p. 550; Am. Elecn., 
Feb., April, Sept., 1904; Mech. Eng., July 25, 1903; Engr., Lond., April 15, 1904; 
Elec. Rev., July 13, 1907. 

75. Effect of Capacity on Efficiency. — In general, as the horse 
power of a boiler increases above normal capacity the over-all efficiency 
will decrease, due to the fact that the furnace and gas passages are 
ordinarily proportioned to effect an evaporation of about 3.5 pounds of 
water from and at 212 degrees F. per square foot of heating surface per 
hour at rated load, the temperature of the escaping gases being from 




BOILERS 



105 



150 to 200 degrees above that of the steam. To increase the rate of 
evaporation more coal must be burned per unit of time and consequently 
a larger volume of gas is generated. The larger the volume of gas the 
higher will be its velocity, which finally reaches a point where heating 
surface is insufficient in extent to absorb the extra heat and as a con- 
sequence the flue gas escapes at a higher temperature, resulting in 
lower boiler and furnace efficiency. With properly proportioned grate, 
furnace and gas passages a boiler may be operated at 100% above stand- 
ard rating with little or no decrease in over-all efficiency. Fig. 40 
shows a case in which the efficiency decreased with the increase in 
capacity, and Fig. 41a illustrates increased efficiency for the higher 
rates of driving. These curves are of value simply as illustrations of 
the behavior in specific cases, and are not applicable to all types of 
boilers. 

TABLE 15. 
RESULTS OF COAL TESTS AT ARMOUR GLUE WORKS, CHICAGO, AUG. 17, 1905. 



Date of Test. 



March 5, 1905 . . 

March 3, 1905 . . 

June 14, 1905 . . 

June 15, 1905 . . 
June 16, 1905 . . 
June 17, 1905 . . 

June 19, 1905 . . 

June 20, 1905 . . 

July 1, 1905. . . . 

July 6, 1905. . . . 

July 28, 1905. . . 

July 29, 1905. . . 

Aug. 5, 1905 . . . 

Aug. 7, 1905 . . . 

Aug. 8, 1905 . . . 

Aug. 9, 1905 . . . 
Aug. 11, 1905 .. 



Name and Kind of Coal. 



Williamson County 
Coal Co.'s, mine run 

Harden & Hafer, mine 
run 

Crerar-Clinch & Co., 
2" screenings. 

...do 

...do 



Brackett Coal and 

Coke Co., lump. 
...do 



do. 



Kelly ville Coal Co. 

mine run. 
Brackett C. & C. Co. 

Keeler mine run. 
Kellyville Coal Co. 

washed pea. 
...do 



Dering Coal Co., mine 
run. 

Dering Coal Co., Sulli- 
van Co., screenings. 

Consolidated Indiana 
Coal Co., Sullivan 
Co., screenings. 

Screenings 



Ziegler, screenings 



Railroad Car 
Number. 



C.C.C.&St.L. 

No. 26368 
S. I. No. 5735 

I.C. No. 88362 

I.C. No. 88362 
I. C. No. 88362 
C. &E. I., No. 

8891. 
C. &E. I. No. 

5002 
C. & E. I. No. 

5002 
C. & E. I. No. 

10030. 
C. & E. I. No. 

12367 
C. & E. I. No. 

6211. 
C. &E. I. No. 

6211 
C. &E. I. No. 

25125 
E. &T. H. No. 

5132. 
E. &T. H. No. 

3239 

E. &T. H. No. 

6534 
I.C. No. 81184 



Cost 
per 
Ton 
Deliv- 
ered. 


Cost to 
Evaporate 

1000 
Pounds of 

Water. 


$1.90 


$0.1531 


1.70 


0.1231 


1.50 


0.1293 


1.50 
1.50 
1.65 


0.1218 
0.1175 
0.122 


1.65 


0.1212 


1.65 


0.1352 


1.595 


0.1355 


1.65 


0.1236 


1.50 


0.1285 


1.50 


0.119 


1.575 


0.125 


1.40 


0.11 


1.35 


0.105 


1.30 


0.0973 


1.50 


0.1047 



Pounds 
Water 

Evapo- 
rated per 

$1.00 of 
Coal. 



6,532 

8,123 

7,734 

8,210 
8,511 
8,197 

8,251 

7,396 

7,380 

8,091 v 

7,782 

8,403 

8,000 

9,091 

9,524 

10,277 
9,551 



* See, "The 
1907, p. 677. 



Nature of True Boiler Efficiency," Jour. Wes. Soc. Engrs., Oct., 



106 



STEAM POWER PLANT ENGINEERING 



In nearly all stations the boilers must have sufficient overload 
capacity to take care of peak loads or to allow some of the boilers to be 
shut down for cleaning or repairs, since the installation of sufficient 



70 


































a 












fy) 






















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D 

> 

a 












=% 


^ 


















































1 

let) 

1 

5 

3 

W 

■o 55 
































5.5 g> 

0) i- 










-.*«$ 






















b.0gc! 








c# 


C^ 






















' 4 - 5 > | 

4.0® 

1 


a 

03 

ft 
































50 
































3.5.Q- 
hi 
























Jour.W.S.E.,Fet 


.1719(4. 





0.45 



0.5 



0.2 0.25 0.3 0.35 0.4 

Draft over Fire in Inches of "Water 
Fig. 40. Influence of Draft on the Efficiency and Capacity of a 350-Horse-power Babcock 
and Wilcox Boiler with Chain Grate. 



12 



3 11 



10 

































































































































— k 


i — , 














































^ 


5\ 






















































































( 


"^X 












































































































































































X 


N K 








































t) 



2 3 4 5 6 

Lb. Water Evaporated per Sq! Ft. of Heating Surface per Hour 

Fig. 41. Effect of Rate of driving on Economy of a 150-Horse-power Stirling Boiler, 

Hand Fired. 



rated boiler capacity would be expensive and in many instances pro- 
hibitive in cost. In small stations, however, too large a boiler capacity 
frequently is to be preferred to an overloaded installation, since the 



BOILERS 



107 



75 


















































« 


• • 

• 






• 
• 


• 


• 




70 


















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• 


• 




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• • 


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S 65 

On 






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• 


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• 


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• • 


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• 










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e eo 

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• 
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• 


• 


• 


























• 


















a 

u 
3 

& 55 

■d 

a 

as 








• / 










500 H.P, B. & W._Boiler 
5000 Sq. Ft. Water Heating Surface 
940 >» " Superheating 
Standard Chain Grate and Setting 
90 Sq. Ft. Grate Surface 






















■2 
o 
«50 
















' various uraaes oi uoai 
Fisk Street Station, Common Wealth-Edison Co. 
Chicago 111. 


































45 


•/ 






• 



























































400 



500 



700 800 900 

Boiler Horse Power 



1000 



1100 



1200 



Fig. 41a. Relation Between Efficiency and Capacity, 500 H.P. Boiler, Fisk Street 
Station, Commonwealth Edison Co., Chicago. 



■ail- 

8 

>.io. 

c 

»10.i 

3 

a 10. 



10, 



§ 10 



2 9.6 

a 

a 

H 9.4 



■3 9.2 

> 

I 9 

















600 


H.P 


B.& W. 


BOILER 














h 










6006 . -■-:-' - - : 

EQUIPPED WITH RONEY STOKERS 

AT THE 59th ST STATION 

OF THE INTERBOROUGH RAPID 

TRANSIT CO., N.Y. 














\ • 
















1 


1 








































• \. 






























1 


sir 


IGLE 


STO 


<ER 


















'% 


^~i 








J 




100 Sq.Ftj.Grate Su 


face 


















•\ 










Coal 14250 B.Tj.U.pt 


xPo 


and 










DOL 
Fr 


BLE 
)nt, 


STOKER N 
100 Sq.Ft. 










Nat 


ural 


Draft; 














Re 


ar, 


80S 


q.Ft. 




"%. 







































600 



760 



920 1000 1080 

Boiler Horse Power 



1160 



1240 1320 



Fig. 41b. Effect of Rate of Driving on the Efficiency of a 600 H.P. B. & W. Boiler. 



108 



STEAM POWER PLANT ENGINEERING 




0.3 0.4 0.5 

Draft La Inches of Water 

Fig. 41c. Influence of Draft on the Capacity of a 600 H.P. B. & W. Boiler. 

TABLE 15a. 

BOILER PERFORMANCE. 

Pounds of Water Evaporated per Hour from and at 212 Deg. F. per pound of Fuel. 



Calorific Value 








Boiler and Furnace Efficiency. 








of Fuel, 






















B.T.U. 






















per Pound. 


40 


45 


50 


55 


60 


65 


70 


75 


80 


85 


7,500 


3.09 


3.48 


3.86 


4.25 


4.64 


5.02 


5.41 


5.80 


6.18 


6.57 


8,000 


3.30 


3.71 


4.12 


4.55 


4.95 


5.36 


5.77 


6.18 


6.60 


7.01 


8,500 


3.51 


3.94 


4.38 


4.81 


5.26 


5.70 


6.14 


6.57 


7.01 


7.45 


9,000 


3.71 


4.18 


4.64 


5.10 


5.56 


6.04 


6.50 


6.96 


7.42 


7.90 


9,500 


3.92 


4.41 


4.90 


5.39 


5.88 


6.47 


6.86 


7.35 


7.85 


8.33 


10,000 


4.12 


4.64 


5.16 


5.66 


6.19 


6.70 


7.21 


7.74 


8.25 


8.76 


10,500 


4.31 


4.86 


5.40 


5.94 


6.48 


7.01 


7.55 


8.10 


8.64 


9.17 


11,000 


4.52 


5.09 


5.65 


6.22 


6.79 


7.35 


7.91 


8.48 


9.05 


9.61 


11,500 


4.74 


5.31 


5.91 


6.50 


7.10 


7.69 


8.28 


8.86 


9.45 


10.0 


12,000 


4.94 


5.55 


6.16 


6.78 


7.40 


8.01 


8.64 


9.25 


9.86 


10.5 


12,500 


5.14 


5.7.8 


6.42 


7.06 


7.70 


8.35 


9.00 


9.64 


10.3 


11.0 


13,000 


5.35 


6.01 


6.69 


7.35 


8.01 


8.69 


9.35 


10.0 


10.7 


11.4 


13,500 


5.56 


6.25 


6.95 


7.65 


8.34 


9.03 


9.72 


10.4 


11.1 


11.8 


14,000 


5.75 


6.48 


7.20 


7.91 


8.64 


9.35 


10.1 


10.8 


11.6 


12.2 


14,500 


5.96 


6.70 


7.45 


8.20 


8.95 


9.70 


10.5 


11.2 


12.0 


12.7 


15,000 


6.18 


6.95 


7.72 


8.50 


9.26 


10.1 


11.8 " 


11.6 


12.4 


13.1 



BOILERS 



109 



extra first cost of the former may be less than the loss due to poor 
efficiency and depreciation in the latter. 

As far as forcing is concerned the fire-tube boiler is as effective as the 
water-tube, more depending upon the furnace, grate surface, draft and 
character of fuel than upon the type of boiler. All boilers are subject 
to more or less priming at heavy overloads, and the overload capacity 
is often limited on this account. 

The Forcing Capacity of Fire-Tube Boilers: F. W. Dean, Trans. A.S.M.E., 26-92. 
Increasing Capacity of Steam Boilers: Kreisinger and Ray, Power, May 24, 1910. 

76. Thickness of Fire. — For a given furnace and boiler, quality and 
size of fuel and intensity of draft, a certain depth of fuel will give maxi- 
mum efficiency. Too thin a fire results in an excess of air and too 
thick a fire in a deficiency, the economy being lowered in either case. 
On account of the number of conditions upon which the proper thick- 
ness depends, it can only be determined for a particular case by actual 
test, the available data being insufficient for drawing conclusions. The 
curves in Fig. 42 are plotted from a series of tests made on a 350-horse- 



300 5 

% 

Q 

200 g 

100 I 
o 





■5 

I 

Is 
H 

- - 

•d Ch 
c 























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ty 


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60 




1 


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Sffi 


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cy 




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■s 














/( 


1 












































































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) 








/ 














































/ 








































30 















































Thickness of Fire, Inches 

Fig. 42. Effect of thickness of Fire on the Capacity and Efficiency of a 350-Horse-power 
Stirling Boiler, equipped with Chain Grate. 



power Stirling boiler equipped with chain grate at the power plant of 
the Armour Institute of Technology. The damper was left wide open 
throughout the test and the speed of the grate kept constant. Ratio 
of grate to heating surface, 1 to 42. Carterville washed coal No. 4 was 
used in all tests. The curves in Fig. 43 refer to the performance of a 
150-horse-power water-tube boiler equipped with chain grate at the 
University of Illinois Engineering Experiment Station at Urbana. 



110 



STEAM POWER PLANT ENGINEERING 



The curves in Fig. 44 are plotted from a series of tests on a 500-horse- 
power Babcock & Wilcox boiler equipped with chain grate at the Fisk 
Street station of the Commonwealth Edison Company, Chicago, 111. In 
these tests the conditions of operation are not exactly comparable, 
but they serve to show the variation of economy with thickness of fire 
in each case. In general, with natural draft, fine sizes of coal necessi- 
tate thin fires, since they pack so closely as to greatly restrict the draft. 
Thin fires require closer attention to prevent holes being burned in 






Horse- Power Developed 
(34 Vz Lbs. of Water Evap. into 
Dry Steam F. & A. 212° per Hr.= l H.P.) 

o o o o o o 




















o c 


L^ 


i/ 




















/ 


^ 












Rate 


d Caps 


tcity o: 


* Boile 


r 


X 


/ 


X 


I 








































*P 


^> 


y^o 




















• 


•^ 


<^ 


f& 


















o ^ 


























m 






















7.5 

c . -1 

H 3 ° 6.5 
a -all 

g g pn 6.0 

§*& 8 

M ** ft 

W 5.5 








n 
























o^^ 
























• 










O 


««&izi. 


2£e__ 


o 


























c 


) « 


» 








• 


• 






6-in. 


Fire 


• 


T 



































15 20 25 30 35 

Dry Coal per Sq. Ft. of Grate Surface per Hr.-Lbs. 



40 



Fig. 43. Effect of Thickness of Fire on the Capacity and Efficiency of a 150-Horse-power 

Water-Tube Boiler. 



spots, and respond less readily to sudden demands for steam, but have 
the advantage of letting the air required pass through the grate, whereas 
thick fires often require air to be supplied above the grate to insure 
complete combustion. Thick fires require less attention and hence are 
preferred by firemen. Where sufficient draft is available thick fires 
are more efficient than thin ones, as the air excess is more readily 
controlled. 



BOILERS 



111 



77. Influence of Initial Temperature on Efficiency. — In general the 
higher the initial temperature of the furnace the greater will be the 
efficiency of the heating surface, since the heat transmitted varies almost 
directly with the difference of temperature between the water and the 
products of combustion. If the heating surface is properly distributed 
so that the final temperature of the escaping gas remains constant, the 
efficiency of the boiler and furnace will increase as the initial temperature 
increases, though not in direct proportion. This is on the assumption 



1000 



». 800 



400 



200 



00 



40 

















































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































' 


' 








^ 


^ 


































































/ 
















* 


^> 




















































































■». 




















































































>> 


«. 














































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































































4 




































B< 


>— 














_i 


?_ 














— « 


j 










































































































































































































































































































































A Boiler 14 Tubes High 
B " 9 " 










































































































































































J 


)U 




VV 


S. 


fcj. 


1.1 




<J 





7 8 9 10 

Thickness of Fire in Inches 



12 



Fig. 44. Effect of Thickness of Fire on the Capacity and Efficiency of a 500-Horse-powei 
Babcock and Wilcox Boiler. 



that the amount of heat generated per hour is the same throughout 
all ranges in temperatures. With a condition where the amount of 
heat generated remains constant and the initial temperature varies, 
the final temperature of the escaping gases remains practically constant, 
and in such cases high initial temperatures are productive of high 
boiler and furnace efficiencies. In practice these conditions are seldom 
realized and high furnace temperatures are not necessarily productive 
of high boiler and furnace efficiencies. Some tests show a decided gain 
in efficiency with the higher furnace temperatures (" Some Perform- 



112 STEAM POWER PLANT ENGINEERING 

ances of Boilers and Chain Grate Stokers, with Suggestions for Improve- 
ments/' A. Bement, Jour. West. Soc. Engrs., February, 1904), and 
others show little if any improvement (" A Review of the United States 
Geological Survey Fuel Tests under Steam Boilers," L. P. Breckenridge, 
Jour. West. Soc. Engrs., June, 1907). The majority of high efficiency 
records, however, are associated with high furnace temperatures. 

78. Cost of Boilers and Settings. — Figures giving the cost of boilers, 
irrespective of type, at so much per horse power are misleading, since 
the cost does not increase in the same ratio as the power. The wide 
variation in cost on the horse -power basis is partly due to the difference 
in rating. For instance, Scotch marine boilers are ordinarily rated at 
8 square feet of heating surface per horse power and return tubular boilers 
at 12 square feet. The price approximates one dollar per square foot of 
heating surface for all boilers over 100 horse power. The cost of water- 
tube and fire-tube boilers may be roughly estimated by the following 
formulas (C. H. Benjamin, Engr. U.S., Nov. 15, 1902): 

(A) Cost in dollars = 500 + 9.2 X rated horse power. (16) 

(B) Cost in dollars = 500 + 8.5 X rated horse power. (17) 

(C) Cost in dollars = 100 + 6.5 X rated horse power. (18) 

(D) Cost in dollars = 100 + 5.0 X rated horse power. (19) 

(A) Horizontal water-tube boilers, 125 pounds pressure, 10 square feet heating 
surface per horse power. 

(B) Vertical water-tube boilers, other conditions as in (A). 

(C) Horizontal return tubular boil ers, 12 square feet heating surface per horse 
power. 

(D) Small vertical fire-tube boiler 

The cost of Scotch-marine boilers rated on a basis of 8 square feet 
per horse power may be estimated by means of formula (A). 

The cost of plain settings may be roughly approximated as follows: 

Horizontal water tube : 

Cost = 400 + 0.8 X rated horse power. (20) 

Return tubular: 

Cost = 300 + 0.7 X rated horse power. (21) 

79. Selection of Type. — Boilers constructed by builders of good 
repute are usually designed for safety, durability, and capacity, and 
rigid specifications and inspection of material and workmanship are 
ordinarily not necessary, as the makers' reputations are sufficient 
guarantee of their worth. Marked departure from standard designs 
must necessarily be specified, but in most cases instructions are limited 
to the extent of heating and grate surface, the character of the furnace, 



BOILERS 113 

and arrangement of setting. Numerous tests on various types of boilers 
show practically the same efficiency provided the furnaces and boilers 
are properly designed, so that the relative merits may be considered 
with reference to (1) durability; (2) accessibility for repairs; (3) facility 
for cleaning and inspection; (4) space requirements; (5) adaptability 
to the type of furnace and stoker desired; (6) capacity; and (7) cost of 
boiler and setting. For high pressure, 150 pounds per square inch or 
more, the water -tube or some form of internally fired boiler in which the 
shell plates are not exposed to the high temperature of the furnace is 
considered safer than the horizontal tubular boiler because the shell 
plates and the seams of the latter must be of considerable thickness in 
large units, and being exposed to the hottest part of the fire are likely 
to give trouble, especially if the water contains scale or sediment-form- 
ing elements. Return tubular and stationary locomotive boilers are 
seldom made in sizes over 200 horse power and hence are not to be 
considered for large units. For sizes over 150 horse power where over- 
head room is limited the return tubular boiler is most commonly 
installed, unless high pressure is essential, in which case the internally 
fired Scotch-marine boiler is peculiarly adaptable. The water -tube 
boiler is usually employed in large central stations for high -pressure 
units of 300 to 500 horse power. The particular type of water -tube 
boiler is to some extent a matter of personal taste on the part of the 
engineer. For small powers and for intermittent operation, small 
vertical or horizontal fire-box boilers have the advantage of low first 
cost. The small air leakage and radiation losses give internally fired 
boilers an advantage over externally fired fire-tube or water-tube types, 
but this is partly offset by the greater extent of regenerative sur- 
face in the setting of the latter.* Internally fired boilers are more ex- 
pensive than the externally fired, though the extra cost of setting and 
foundation in the latter may bring the total cost of the entire equip- 
ment to practically the same figure. The design and installation of the 
boilers and furnaces should be left at the outset to a capable engineer. 

Makers usually request the following information from intending pur- 
chasers: i. Steam pressure desired. 

2. The quantity of steam demanded. 

3. The kind of fuel to be burned. 

4. The type of furnace or stoker. 

5. The nature and intensity of draft. 

6. Nature of setting. 

7. Probable temperature of the feed water. 

* At the power plant of the Cosmopolitan Electric Co., Chicago, the brick settings 
of the boilers (500 h.p. B. & W.) are completely encased with riveted boiler plates. 



114 



STEAM POWER PLANT ENGINEERING 



The complete specifications for a return tubular boiler are given in 
Chapter XIX. 

Choice of Boilers for Steam Power Plants : Am. Elecn., June, 1900, p. 261, Dec, 
1905, p. 633, May, 1901, p. 217, June, 1903, p. 256; Engr., Lond., Sept., 1898, p. 232, 
May 29, 1903, p. 555; St. Ry. Rev., Oct., 1901, p. 789; Elec. Rev., Lond., June 16, 
1899, p. 973; St. Ry. Jour., Oct. 5, 1901, p. 453, Oct., 1905, p. 731; Mech. Eng., 
May 16, 1903; Min. Rept., Feb. 21, 1907. 

80. Grates. — Grates may be divided into three general classes, 
namely, stationary, rocking, and traveling grates. The latter are 
treated in Chapter IV. Stationary grates are generally made of cast- 
iron sections in a variety of shapes as illustrated in Fig. 45. The 
bars are ordinarily from 3 inches to 4 inches deep at the center (this 
makes them strong enough to carry the load caused by the weight of 
the fuel without sagging even when the top is red hot), } inch wide at 
the top, and taper to | inch at the bottom to enable the ashes to drop 
clear. The width of the air space is determined by the size of the fuel 
to be used, the average proportions being given in Table 16. 



TABLE 16. 

AIR SPACES AND THICKNESS OF GRATE BARS. 



Size and Kind of Coal. 



Thickness of 
Grate Bars. 



Screenings 

Anthracite — 

Average 

Buckwheat 

Pea or nut 

Stove 

Egg 

Broken 

Lump 

Bituminous, average 
Wood — 

Slabs 

Sawdust 

Shavings 




The " Tupper " and " Herringbone " grate bars are stiffer and less 
likely to warp than the common form, but are not so readily sliced and 
therefore not so convenient with coal that clinkers badly. Sawdust 
or pin-hole grates are used in burning sawdust, tan bark, and very small 
sizes of coal. Grates are often set horizontally and the bars are held 
in place simply by their own weight, but long grates are best placed 
sloping toward the rear to facilitate firing. The front of the grate when 



BOILERS 



115 



designed for bituminous coal is often made solid, this portion being 
called the " dead plate." It serves to hold the green fuel until the 
hydrocarbons have been distilled off, when the charge is pushed back 
on the open grate at the time of next firing. The length of a single bar 
or casting should not exceed three feet. The length of grate may be 



COMMON BAR 




TUPPER 



HERRINGBONE 



K 



• ••• 
»• • • 

• • •• 



SAW- DUST 

Fig. 45. Types of Grate Bars. 



made of two or three bars and should not exceed 6 feet with bituminous 
coal, as this is the greatest length of fire that can be readily worked by 
a stoker. With buckwheat anthracite furnaces 12 feet in depth are not 
unusual, as anthracite fires require no slicing. 

The disadvantage of using stationary grates is that the fire is not 
easily cleaned. Unless the air spaces are kept free of clinkers and 
ashes, combustion is hindered and the fire rendered sluggish. Frequent 
cleaning, however, is wasteful of fuel and reduces the furnace efficiency 
by letting in a large excess of air every time the fire door is opened. 

81. Rocking Grates. — Shaking grates have the advantage of per- 
mitting stoking without opening the fire door and require less manual 
labor than stationary grates. There is a great variety of sectional 
shaking grates on the market and some of them are made self-dumping. 
One of the best known types is illustrated in Fig. 46. Each row or 
section of grate bars is divided into a front and a rear series by twin 
stub levers and connecting rods. An operating handle is adapted to 
manipulate either one or both of the levers in such a manner that the 
front and rear series may operate separately or together. The shaking 
movement causes no increase in the size of the openings and hence pre- 
vents the waste of fine fuel. Ordinarily the width of the grate is made 



116 



STEAM POWER PLANT ENGINEERING 



equal to two or more rows of grate bars so that the live fire may be 
shoved sidewise from one row to the other when cleaning. A depth of 




V V V V I I < V V V V V T^? 



p* 



Fig. 46. A Typical Rocking Grate. 

fire of from 6 to 10 inches is carried according to the nature of the fuel 
and the available draft. 

GfateBars: Engr. U.S., Nov. 1, 1906, p. 728, Jan. 1, 1907, p. 68; Am. Elecn., 
Jan., 1904, p. 269. 

82. Blow-Offs. — Boilers must be provided with blow-off pipes for 
draining off the water and for discharging sediment- and scale -forming 
material. The " bottom blow " is ordinarily an extra heavy pipe of 
suitable diameter connected to the lowest part of the boiler and fitted 




Fig. 47. Horizontal Blow-off Connection 
to Head. 



Fig. 48. Vertical Blow-off Connection 
to Shell. 



with a valve or cock, or both. (See paragraph 349.) Fig. 47 shows an 
arrangement of horizontal blow-off connected to the head of a return 
tubular boiler and Fig. 48 a vertical blow-off connected to the shell. 






BOILERS 



117 



The latter is the better arrangement. The blow-off pipe where it 
passes through the back connection is covered with magnesia, asbestos, 
or fire brick. When exposed to the action of extremely hot furnace 
gases as in forced draft installations, the arrangement illustrated in Fig. 
49 is sometimes used to prevent the pipe from burning out. When the 
blow-off cock is shut and the valve on the vertical branch open there 
is a continuous circulation of water. Where boilers are arranged in 
batteries, the battery may have a common outlet for the blow-off 



_...__ WATER LEVEL _ 



^ 



! i 




Fig. 



49. Blow-off Connection with 
Circulating Pipes. 



Fig. 



50. Blow-off Tank and 
Connections. 



pipes, as illustrated in Fig. 388. Usually the blow-off pipes may dis- 
charge into the open air, but this is not permissible in large cities, nor 
is it lawful to blow directly into the sewer. In this case the water 
and sediment may be discharged into a blow-off tank and permitted to 
cool before delivery to the sewer, as illustrated in Fig. 50. 



WATER LEVEL 




Fig. 51. Surface Blow-off. 



" Surface blows " are often installed to remove scum, grease, and float- 
ing or suspended particles of dirt. The bell -mouthed shape shown in 
Fig. 51 permits the skimmer to accommodate itself to varying water 



118 



STEAM POWER PLANT ENGINEERING 



level, and it is sometimes provided with a float and with a flexible joint, 
Fig. 52. 

83. Damper Regulators. — For maximum furnace efficiency the 
draft must be regulated to burn just enough fuel to supply the steam 



s-— i 




Fig. 52. Buckeye Skimmer. 

required. Where forced draft is employed this is done by regulating 
the speed of the blower. With natural draft it is the usual practice to 
regulate the draft by means of dampers placed in the uptake, and in 
order that the regulation may be effective it should be automatic. 
Automatic dampers are economical and useful and are particularly 
desirable in plants where the demand for steam fluctuates rapidly. 
There are several successful types on the market, some operated by 
water pressure and others by direct boiler pressure. Fig. 53 illus- 




Kitts Hydraulic Damper Regulator. 



trates such a mechanism. Full boiler pressure acting at all times on a 
diaphragm A raises or lowers a weight W attached to arm D according 
as the steam pressure increases or decreases. Arm D actuates a small 
valve V which controls the supply of water to chamber B. A diaphragm 
in chamber B raises and lowers the damper as the water pressure varies, 
a drop of 0.5 pound being sufficient to open the damper to its maximum. 



BOILERS 



119 



The steam diaphragm has a movement of only 0.01 inch and the 
water diaphragm 0.5 inch. When properly adjusted and given proper 
attention automatic dampers work in a very satisfactory manner. 

Fig. 54 shows a section through the Tilden damper regulator, illus- 
trating the principles of the steam-actuated type. The device is con- 





Fig. 54. Tilden Steam Actuated 
Damper Regulator. 



WATER 

Fig. 55. Simple Water Column. 



nected directly to the boiler by pipe A. The pressure on piston B is 
balanced by spring C under normal conditions of operation. Any 
variation from the normal will cause the rod R to move up or down, so 
that the dampers are opened or closed in proportion to the change in 
pressure. The chamber N is separate from chamber M, so that steam 
or water cannot come in contact with the spring. Piston D acts as a 
guide only. 

Damper Regulators : Engr. U.S., Jan. 1, 1907, p. 58; Elec. Wld., May 2, 1908. 

84. Water Gauge. — The water level in a boiler is usually indicated 
either by a gauge glass, by try cocks, or both, connected directly to the 
boiler as in Fig. 1, or to a water column or combination as in Fig. 55. 



120 



STEAM POWER PLANT ENGINEERING 



Each gauge-glass connection should be fitted with a stop valve which 
may be closed in case the tube breaks. In large boilers these valves, 
usually of the quick-closing type, are conveniently operated from the 
boiler-room level by means of a chain attached to the valve stem. 
Self-closing automatic valves are frequently employed, one type being 
illustrated in Fig. 56. If the glass breaks the outrush of steam forces the 
ball against a conical seat and shuts off the supply. When a new glass 
is inserted the ball is forced back by slowly screwing in the valve stem. 
Hinged valves mechanically operated from without are considered 
more reliable than ball valves, as scale is less likely to render them 
inoperative. 




r\ 



BALL VALVE 



<^ 




Fig. 56. Water Guage with Self-closing 
Valve. 



WATER 

Fig. 57. Combined Water Column 
and High and Low- Water Alarm. 



Try cocks or gauge cocks are set at points above and below the desired 
water level, preferably connected directly to the boiler shell but some- 
times to a water column as in Fig. 55. The water level is ascertained 
by opening the cocks in succession. 

The objection to the latter arrangement is that accident to or a 
stoppage of the piping renders both gauge glass and try cocks useless. 
Water columns should be blown out once a day and the gauge cocks 
opened to see that the height of the water indicated tallies with that 
shown by the glass. Some engineers prefer two separate columns to 
each boiler and no cocks, others rely solely upon cocks. 

The water column shown in Fig. 57 has an alarm whistle, controlled 
by two floats, which gives a high and low-water alarm. Numerous 
devices of this class are on the market but they are usually regarded as 






BOILERS 



121 



unreliable and most engineers are content to depend upon water gauge 
and try cocks. 

Water Gauges and Columns: Mach., Sept., 1905, p. 31; Power, Aug., 1905, p. 483; 
Am. Elecn., July, 1904, p. 359; Engr. U.S., Jan. 1, 1907, p. 58. 

85. Fusible or Safety Plugs. — Fusible or safety plugs as illustrated 
in Fig. 58 are brass plugs provided with a fusible metal core. They 



INSIDE TYPES 



OUT6IDE,TYt»ES 




Fig. 58. Types of Fusible Plugs. 

are inserted in the shell or tubes at the lowest permissible water line. 
When covered by water the heat is conducted away sufficiently fast to 
keep the temperature below the fusing point, but when uncovered the 
low conductivity of the steam prevents the rapid withdrawal of heat, 
whereupon the alloy melts and the blast of escaping steam gives warning. 
The melting point of fusible metals being sometimes uncertain, plugs 
occasionally blow out without apparent cause and at other times fail to 
act when shell is overheated. Fusible plugs are required by law in 
many cities. 

86. Mechanical Tube Cleaners. — Although purifying plants, boiler 
compounds, and the like are preventive of scale formation to a great 
extent, experience shows that the most satisfactory method is to use 
mechanical tube cleaners for cutting or breaking the scale. The prin- 
ciples of construction of these devices vary widely according to the types 
of boilers in which they are used, and depend upon the nature of the 




C D F t 

Fig. 59. Mechanical Tube Cleaner, 



Hammer Type. 



duty which they must perform. They may be conveniently divided into 
two classes: 

1 . Those which loosen the scale by a series of rapid hammer blows, 
Fig. 59. 

2. Those which cut out the scale by a revolving tool, Fig. 60. 



122 



STEAM POWER PLANT ENGINEERING 



The hammer device is applicable to either the water- or fire-tube type 
of boiler, but the revolving cutter is applicable to the water-tube only. 
Steam, compressed air, or water under pressure may be used as the motive 
power, though the latter is the most convenient and satisfactory. 

Referring to Fig. 59, the hammer head J is given a rapid motion, 
which may reach 1,500 vibrations per minute, and subjects the tube 
to repeated shocks, thereby cracking the brittle scale and jarring it 
loose from the water surface of the tube. The cleaner head is attached 
to a flexible pipe of sufficient length to enable it to be pushed from one 
end to the other. Even if carefully manipulated the hammer is apt to 
injure the tube by swaging it to a larger diameter, producing crystalliza- 
tion in the metal and causing leaks where the tubes are expanded into 
the sheets, hence its use is not to be recommended. 

Hydraulic turbine cutters are made in many designs, one of which 
is shown in Fig. 60. The cylindrical casing D contains a hydraulic 




£z£Z$zz£SZgB 



i I 



aaajajakaatz 



Fig. 60. Mechanical Tube Cleaner, — Turbine Type. 

turbine consisting of a fixed guide plate which directs the water at the 
proper angles upon the vanes of the turbine wheel T. The cutters C 
revolve at high speed and chip the scale into small pieces. The stream of 
water flowing from the turbine envelops the cutters, keeps their edges 
cool, and washes away the scale as fast as it is detached. Different 
styles of cutter wheels are furnished with each cleaner so as to adapt 
the device to all kinds of scale formations. In well-managed plants 
scale is not permitted to deposit to a thickness greater than & to & of 
an inch. 

The soot and cinders which accumulate on the inside surface of fire- 
tube boilers are removed by mechanical scrapers, brushes, or steam-jet 
blowers. (For a description of these devices see American Electrician, 
April, 1904, p. 576.) 

The tubes of a water -tube boiler are cleaned externally by means of 
a steam jet. 

Boiler Accessories : Am. Elecn., April, 1903, p. 194, Feb., 1905, p. 67, June, 1904, 
p. 269, July, 1904, p. 339; Am. Mach., April 21, 1901, p. 518; Engr. U.S., Jan., 1907, 
p. 56. 

Boiler Arches : Boiler Maker, Aug., 1907. 






BOILERS 123 

Blow-off Connections : Locomotive, Oct., 1906; Elec. World, Nov. 2, 1907; Nat. 
Engr., June, 1904; Eng. Rec, May 9, 1908. 

Bracing: Boiler Maker, April, 1905; Mach., Sept., 1903, p. 18, Oct., 1903, p. 83; 
Power, Jan., 1903, p. 24, Oct., 1905, p. 611, Nov., p. 687; Eng. News, Dec. 15, 1904, 
p. 533; Trans. A.S.M.E., 18-989; Am. Mach., June 3, 1897, June 2, 1898, p. 404; 
Engr. U.S., Jan. 1, 1907, p. 18; Engr., Lond., April 25, 1900, pp. 412, 419; Prac. 
Engr., Jan., 1907. 

Boiler Cleaning : Am. Elecn., Dec, 1900, April, 1904, p. 174; Power, May and Oct., 
1905, Aug., 1906, p. 465; Locomotive, Oct., 1904; Boiler Maker, Aug., 1905; Engr. 
U.S., Jan. 1, 1907, p. 109. 

Boiler Design: Engr. U.S., Jan. 15, 1902, p. 59; Eng. Mag., May, 1904, p. 233; 
Eng. Rec, July 14, 1900, May 18, 1901, p. 467, Oct. 12, 1901, p. 347; Power, Oct., 
1901, p. 14, March, 1906, p. 147; St. Ry. Rev., Feb. 15, 1899, p. 125; West. Elecn. # 
April 20, 1901, p. 267; Am. Mach., April 21, 1904; Mach., Oct., 1902. 

Boiler Dimensions : All Types of Stationary Boilers : Eng. U.S., Jan. 1, 1907, p. 10, 
Aug. 1, 15, 22, 1903. Small Marine Boilers : Am. Mach., Sept. 3, 1896, p. 823. 
Tubular Boilers : Mach., Oct., 1902, p. 94. 

Circulation in Boilers: Eng. Rec, July 20, 1901; Cassier's Mag., Jan., 1905; Elec. 
Rev., Lond., April 4, 1902; Engng., April 18, 1902; Engr., Lond., Nov. 6, 1903; Am. 
Mach., Jan. 14, 1897, p. 40, Sept. 20, 1900, p. 910; Eng. News, Jan. 18, 1900, p. 40; 
Trans. A.S.M.E., 7-814, 9-489; Engr. U.S., Oct. 15, 1907. 

Domes : Engr. U.S., Jan. 1, 1907, p. 27. 

Classification of Boilers and Comparison of Types : Engr. U.S., Jan. 1, 1907; Min. 
Rept., Feb. 21, 1907. 

Furnace and Settings: National Engr., Sept. and Nov., 1907; Elec. World., Sept. 
7, 1907; Am. Elecn., Jan., 1902, p. 10, Nov., 1903, p. 557, July, 1904, p. 339; Trans. 
A.S.M.E., 6-118, 16-590, 19-74, 782, 20-95; Am. Mach., Aug. 18, 1898; Engr. U.S., 
July 15, 1905, p. 471, Sept. 15, 1905, p. 622, Aug. 1, 1906, p. 491, May 15, 1906, Jan. 
1, 1907; Power, March 24, 1908, p. 445, June, 1905; Eng. Mag., July, 1897, p. 587. 

Boiler Inspection: Power, Jan., 1906, p. 32; Engr. U.S., Feb. 15, 1907; Trans. 
A.S.M.E., 4-142; Boiler Maker, Nov., 1907; Cassier's, Feb., 1907. 

Riveted Joints : Power, March, 1906, p. 147, April, 1906, p. 227; Engr. U.S., Jan. 1, 
1907, p. 21, Aug. 15, 1907, p. 784; Trans. A.S.M.E., 6-120, 10-707; Boiler Maker, 
June, 1906, Dec, 1907; Prac Engr., Dec 13, 1907. 

Safety Valves : See paragraph 350. 

Specifications: Power, Dec, 1905, p. 728; Nat. Engr., May 15, 1903, p. 367; 
Eng. News, Oct. 20, 1898,' p. 251; Boiler Maker, Sept., 1906, p. 243. 

Thickness of Boiler Plate : Am. Mach., Jan. 16 and Feb. 27, 1902; Trans. A.S.M.E., 
22-127, 15-629, 24-921; Eng. News, Jan. 31, 1901, p. 121. 

Boiler Testing : See A.S.M.E. Code for conducting Standard Boiler Trials, reprinted 
in Appendix B. See also Power and Engr., Feb. 23, 1909. 

Bridge Walls in Theory and Practice: Power and Engr., Mar. 9, 1909, p. 452. 



CHAPTER IV. 

SMOKE PREVENTION, FURNACES, STOKERS. 

87. General. — It is recognized that bituminous coal can be 
efficiently burned without smoke in a properly designed furnace if 
proper attention is given to stoking and other factors involved. It is 
nevertheless a common statement among owners of power plants that 
it is cheaper to smoke than to operate without smoke. This is 
^undoubtedly true in many cases where smokeless combustion can be 
secured only by admitting a considerable excess of air with a consequent 
loss in economy frequently greater than that due to incomplete com- 
bustion and smoke. Even under the most favorable conditions, how- 
ever, smokeless combustion depends largely upon skillful manipulation 
by an interested and efficient fireman.* The order of intelligence 
demanded for this work is out of all proportion to the wages paid. In 
many small plants — and these are usually the most obstinate smoke 
offenders — the fireman handles as much as a ton of coal per hour by 
hand, besides caring for the feed pumps and water levels, keeping the 
boilers clean, and removing the ash. The boiler room is frequently 
poorly lighted and poorly ventilated. It is, therefore, not surprising 
that the fireman seldom worries about the smoke problem. A better 
wage scale and more consideration for the firemen might do a great deal 
toward abating the smoke nuisance. In order that combustion may be 
smokeless and efficient, the volatile gases and separated free carbon must 
be brought into intimate contact with the proper quantity of air and main- 
tained at a temperature above the ignition point until oxidation is complete 
before they are brought in contact with the heat-absorbing surfaces of the 
boiler. Mere excess of air will not effect smokeless combustion, even if 
the gases and air are thoroughly mixed, if the temperature is prema- 
turely reduced below that necessary for combustion by contact with the 
heat-absorbing surfaces of the boiler. 

Smoke may be produced, therefore, by 

1. An insufficient amount of air for the perfect combustion of the 
volatile gases. This is primarily a question of draft. 

2. An imperfect mixture of air and combustible. 

3. A temperature too low to permit complete oxidation of the 
volatile combustible. 

* See Appendix G. Rules for firemen using Illinois and Indiana coal in hand- 
fired furnaces. 

124 






SMOKE PREVENTION, FURNACES, STOKERS 125 

Smoke-preventing devices may be divided into two classes: 

(1) Those which are an integral part of the boiler and setting, such 
as mechanical stokers, Dutch ovens, down-draft furnaces and fire-tile 
combustion chambers incorporated with the regular setting. 

(2) Those which may be conveniently attached to plants already in 
Operation without material modification of the furnace, such as steam 
jets and other means of mixing air and combustible gases, admission of 
air through the bridge wall or side wall, and mechanical draft. 

88. Mechanical Stokers. — Uniform evolution of the volatile gases of 
the fuel is the essential requisite for smokeless combustion, and it is 
for this reason that mechanical stokers as a class are more effective in 
preventing smoke than any apparatus accompanied by intermittent 
firing. Stokers which feed irregularly have the effect of hand-fired 
furnaces, and it is necessary not only to employ some powerful auxiliary 
mixing device but also to furnish at times an extra supply to take care 
of the enormous volume of volatile gas evolved after a fresh charge of 
fuel is added. 

Carefully adjusted automatic stokers owe their high efficiency to 
(1) uniformity of feed; (2) proper proportion of air and combustible; (3) 
absence of excessive air dilution, as when the fire doors are opened in 
connection with hand firing; and (4) self-cleaning grates. 

Daily records are essential with any type of stoker or hand firing if 
efficient results are expected, as only by frequent observation isit possible 
to determine the proper adjustment of air supply, depth of fire, rate of 
feed, and the like. Control of air supply is almost as important as the 
upkeep and effective operation. In the best firing practice the right 
amount of air, depth of fire, and rate of feed must be worked out by the 
engineer. 

Stokers are often condemned by owners as inefficient and inferior to 
hand stoking because no particular attention had been paid to them 
beyond filling the hopper with coal. They should be operated in strict 
accordance with the principles of design. 

In plants of 2000 horse power or over the installation of mechanical 
stokers and coal conveyors effects a considerable saving of labor and 
can usually be relied upon to solve the smoke problem if reasonable 
attention is given to their operation. In smaller plants interest on 
the investment and other considerations may make hand firing more 
economical, although many plants of capacities as small as 200 horse 
power are giving satisfaction, particularly in places where a poor grade 
of fuel is used and smoke ordinances are rigidly enforced. A stoker 
of the self-cleaning, slow-running type requires much less attention 
than the hand-fired furnace. With hand firing one fireman can effi- 



126 STEAM POWER PLANT ENGINEERING 

ciently attend to the water, coal, and ashes of about 200 horse power 
or handle coal for say 500 horse power, whereas with good automatic 
stokers he can readily take care of 2000 horse power or of 4000 horse 
power with chain grates equipped with overhead bunkers and down 
spouts. 

The best stokers are those which are least complicated and simplest 
in operation. A cheap stoker is a poor investment, since the cost of 
repairs and shut downs will usually amount to more than the saving 
in price. 

The following outline gives a classification of a few of the best known 
American mechanical stokers: 

Front Feed. 

Chain Grates: Step Grates: 
Babcock & Wilcox, Roney, 

Green, Wilkinson, 

McKenzie, Acme, 

Playford. McClave. 

Side Feed. Under Feed. 
Step Grates: Jones, 

Murphy, American, 

Detroit, Taylor, 

Mode. Guckett. 

Down Draft. Sprinkler. 

Hawley Little Giant, 

Swift, 
Powdered Fuel. Vulcan. 

See paragraphs 34-46. 



Mechanical Stokers: Power, March, 1906, p. 189, Aug., 1905, p. 487, March, 1903, 
p. 112 ; Jour. West. Soc. Engrs., Feb., 1903, p. 44; Eng. Soc. West Penn., April, 1903, 
p. 169 ; Eng. News, March 26, 1903, p. 272 ; Eng. News, 35-226 ; Eng. Mag., July, 1902, 
p. 528; Engr. U.S., Jan. 1, 1907, p. 83, Aug. 15, 1906, p. 440, July 2, 1906, p. 437, 
Feb. 1, 1904, p. 114, April 1, 1903, p. 262 ; Elec. Engr., Lond., 33-977 ; Cassier's, Sept., 
1906, p. 469; Trans. A.S.M.E., 17-278, 558; Am. Engr., July, 1905, p. 281, July, 
1904, p. 284; Am. Elecn., July, 1904, p. 329, 1902, p. 489, 14-18, 12-411, 263. 

89. Chain Grates. — The chain grate, Fig. 61, is one of the most 
popular forms of automatic stokers. It embodies a moving endless 
chain of grate bars mounted on a frame with provision for the con- 
tinuous and uniform feeding of coal into the furnace, the fuel and the 
grate moving together. The operations of feeding the coal, carrying 
it through the progressive stages of combustion, removing the ashes and 
clinkers, and maintaining a clean grate and free air supply are practi- 



SMOKE PREVENTION, FURNACES, STOKERS 



127 




128 



STEAM POWER PLANT ENGINEERING 




Fig. 62. Babcock and Wilcox Boiler, Chain Grate, Ordinary Setting. 




Fig. 63. Babcock and Wilcox Boiler, Chain Grate, Fire-tile Roof. 



SMOKE PREVENTION, FURNACES, STOKERS 



129 



cally automatic. The driving mechanism consists of a gear train 
actuated by a cast-steel ratchet and pawls, the arms carrying the latter 
being given a reciprocating motion by an eccentric mounted on a line 
shaft. The latter may be driven by any type of engine or motor and 
the speed of the grate regulated by varying the stroke of the arm carry- 
ing the pawls. Fuel is fed into a hopper placed at the front end of the 
furnace and the depth of the fuel regulated by a guillotine damper. 
The entire grate and driving mechanism are mounted on a permanent 




Fig. 64. Section of Tiles Encircling Lower Row of Tubes. 

truck and may readily be removed from beneath the boiler. The front 
part of the furnace is provided with a flat or slightly inclined ignition 
arch as illustrated. The thickness of the fire and the speed of the grate 
should be so regulated that when the fuel has reached the end of the 
grate it shall have been completely consumed and ashes only will be dis- 
charged into the pit. The combination of chain grate with inclined igni- 




Fig. 65. Tiles Between Lower Row of Tubes, Back of Encircling Tiles. 



tion arch, curved bridge wall, and lower course of tubes covered with fire 
tiling, as shown in Fig. 61, makes an excellent smokeless furnace, though 
the depth of fire and air supply must be carefully regulated to prevent 
excessive air dilution. With chain-grate stokers there may be a con- 
siderable leakage of air between the grate and bridge wall, through the 
coal in the hoppers, under the coal gate, and through the fire bed at 
the rear where it is mostly ash. Various schemes have been employed 
to prevent leakage at the end of the grate by using water backs, ash-pit 
dampers, and the like. Fig. 62 shows the application of the chain grate 



130 STEAM POWER PLANT ENGINEERING 

to a Babcock & Wilcox boiler with the standard arrangement of fire 
tiles and bridge wall. To insure the best coking effect with this ar- 
rangement the flat ignition arch is employed. This reduces the velocity 
and increases the temperature of the gases as they are distilled from 
the green fuel, thereby assisting complete combustion. Under normal 
conditions of operation this insures practically smokeless combustion, 
but is not always successful at heavy overloads, since the length of 
furnace is not always great enough to thoroughly mix the air and 
combustible gases before they reach the boiler surfaces. A more 
satisfactory arrangement is the modification illustrated in Fig. 63. The 
coking arch is slightly inclined and the gases are compelled by means 
of the fire-tile roof to take a longer path to the rear before crossing 
the tubes to the uptake, which must necessarily be at the front end 
of the boiler. Tests have shown this arrangement to give excellent 
efficiency of boiler and furnace, with practically smokeless combustion, 
though the cost of the upkeep of the tiling is very high. 

Figs. 64 and 65 illustrate the method employed in suspending the 
fire tiles. Another method of increasing the extent of regenerative 
fire-brick surface without disturbing the standard setting is to cover 
the tubes over the furnace with a fire tile, which permits the gases to 
flow between the tubes. Such an application to a Stirling boiler is 
illustrated in Figs. 66 and 67. The useful life of this class of tiling is 
very short and this fact prevents its adoption in most cases. 

Fig. 67a and Fig. 67b show recent applications of chain grates to 
large B. & W. boilers which are effecting high efficiency and heavy over- 
load capacity with practically smokeless combustion. See Power and 
The Engineer, May 31, 1910, p. 981. 

Chain grates properly installed and taken care of cost very little for 
maintenance. At the South Side Elevated Railroad Station, Chicago, 
where there are 22 B. & W. chain grates, the total cost of repairs for 
the grates alone is stated to have been but $25 in eight years. Fiat 
fire-brick arches as in Fig. 62 have to be renewed about once a year. 
The inclined arch with a good quality of fire brick should last two years 
or more. The cost of renewing arches of this type approximates $200 
for 500-horse-power boilers per setting. The fire-tile furnace is more 
costly in yearly repairs than the common furnace, but the increased 
efficiency may offset the extra cost. (A. Bement, Eng. U. S., June 1, 
1907, p. 606.) The repairs for a 500-horse-power Babcock & Wilcox 
boiler setting as in Fig. 63 is approximately $350 per year for arches 
and tiling. 

90. Step Grates, Front Feed. — Fig. 68 shows the general arrange- 
ment of a Roney stoker and Fig. 69 that of a Wilkinson stoker, illus- 



SMOKE PREVENTION, FURNACES, STOKERS 



131 




Fig. 66. Application of "Economy" Fire Tiles to Stirling Boiler. 



SECTION THROUGH A-B 



p r=i 


r^\ ' r°l 




— L . L 


M 


— i — r 


pJL 


-J. p-1 
i — r 




_ L ^ 
— V ' — } 


Ud kj 


^ V- 



Fig. 67. Method of Anchoring "Economy" Fire Tiles to Tubes. 



132 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



138 




134 



STEAM POWER PLANT ENGINEERING 



trating the step-grate, front-feed principle. The Roney stoker consists 
of a hopper for receiving the coal, a set of rocking stepped grates inclined 
at a proper angle from the horizontal, and a dumping grate at the bottom 
of the incline for receiving and discharging the ash and clinkers. The 
dumping grate is divided into several sections for convenience in han- 
dling. The coal is fed on to the inclined grate from the hopper by a 
reciprocating " pusher " actuated by the " agitator." The power is 
supplied through an eccentric operated by a small engine or motor. 
The normal feed is about 10 strokes per minute. The grate bars rock 




Details of Construction of the Roney Mechanical Stoker 
Fig. 68. Details of Ronev Stoker. 



through an arc of 30 degrees, assuming alternately horizontal and 
inclined positions. The construction permits abundance of air to pass 
through the fuel, with little or no possibility of coal dropping through 
the grate. A coking arch of fire brick is sprung across the furnace as 
indicated. This stoker operates with natural draft and with suitable 
arrangement of fire tiling effects complete and efficient combustion. 
Without a fire tile roof construction smokeless combustion is effected 
with difficulty, particularly at heavy loads. 

In the Wilkinson stoker the inclined grate bars are hollow and are 
arranged side by side, every alternate bar being movable. When in 



SMOKE PREVENTION, FURNACES, STOKERS 



135 




Fig. 68a. Double Stoker Installation at the 59th Street Station of the 
Interborough Rapid Transit Co., N. Y. 



136 



STEAM POWER PLANT ENGINEERING 



operation there is a constant sawing action of the grate bars, causing 
the fuel to flow forward and downward. A small steam jet with about 
T V inch opening is introduced into the end of each hollow grate bar, 
thus inducing the required amount of air for combustion, which 
passes through air openings approximately \ inch wide by 3 inches 



THE MECHANISM OF THE WILKINSON STOKER. 





Fig. 69. Details of Wilkinson Stoker. 



long. These stokers are driven by two small hydraulic motors, the 
water being furnished by a small pump and being used over and over 
again. 

91. Step Grates, Side Feed. — Fig. 70 shows a front vertical section 
and Fig. 71 a side vertical section through a Murphy automatic stoker 
and furnace. The apparatus is in effect a Dutch oven equipped with 
an automatic feeding and stoking device. Coal is introduced either 
mechanically or by hand into the magazine at each side of the furnace 
and above the grate and descends by gravity upon the coking plate. 
Reciprocating stoker boxes push the coal out upon the grate bars. 
Every alternate grate bar is movable and pivoted at its upper end. 
A rocker bar driven by a small motor or engine causes the lower ends to 
move up and down, this action producing the required stoking effect. 
A device for grinding up the clinker and ash is provided as shown at 
the bottom of the furnace. This is hollow and is connected by a 2-inch 






SMOKE PREVENTION, FURNACES, STOKERS 



137 




TMNSVERSE SECTION. 
Fig. 70. Murphy Furnace, Front Section. 




Fig. 71. Murphy Furnace, Side Section. 



138 



STEAM POWER PLANT ENGINEERING 



pipe with the smoke flue, so that the cold air passing through it prevents 
it being destroyed by the heat. Air is supplied to the green coal through 
flues passing under the coking plates, and the speed of the stoker boxes 
and grate bars can be regulated to conform to any rate of combustion. 
On account of the large fire-brick combustion chamber, this stoker with 
careful manipulationis capable of practically smokeless combustion. The 
power house of the Northwestern Elevated Railroad Company, Chicago, 
111., is equipped with Murphy furnaces, which are operating smoke- 
lessly at an unusually high combustion rate, whereas a number of 
other installations using the same type of stoker and boiler and 
burning the same class of fuel are heavy smoke producers. Murphy 
furnaces are peculiarly adapted to variable loads, since at light loads 
the stoker may be operated with reduced grate area by allowing the 
bottom of the grate to partly fill with ashes. 

92. Underfeed Stokers. — Fig. 72 shows the general principles of the 
Jones underfeed stoker. It consists of a steam-actuated ram with a 




Fig. 72. Jones Underfeed Stoker. 

fuel hopper outside the furnace proper and a retort or fuel magazine 
and auxiliary ram, A, A, within. Heavy cast-iron tuyere blocks for 
the admission of air are placed on either side of the retort. Fuel is 
forced underneath the fire by the ram and its auxiliary, the ram 
movement of the fuel being backward and upward, displacing the incan- 
descent fuel and ash and forcing it on to the dead plates. There is 
no ash pit, the ashes being raked from the dead plate by hand. Air, 
supplied by a blower, is admitted through the openings in the tuyere 
blocks. The latter are at a point above the green fuel in the retort 
tube, but below the fire. The standard size of retort is about 6 feet 
in length by 24 inches in width and 18 inches in depth, and experience 
has shown that other sizes are not necessary, since the spaces between 
retort and side walls of the various furnaces may be provided for by 
extending the width of dead plate. One stoker is usually installed for 
each furnace, though two are sometimes required. The draft from 
the fan and the number of strokes of the ram are automatically con- 



SMOKE PREVENTION, FURNACES, STOKERS 



139 



trolled by the steam pressure, although provision is made for regulating 
either by hand. Underfeed stokers are adaptable to all grades and sizes 
of coal and on account of the forced draft are capable of burning very 
low grades of fuel. A number of these stokers are installed in the 
power plant of the First National Bank Building, Chicago, and are 
giving high efficiency and smokeless combustion with low-grade Illinois 
screenings. The cost of upkeep is rather high when compared with 
chain grates. 

Fig. 73 shows an application of an American underfeed stoker to a 
return tubular boiler. This differs from the Jones stoker in the method 





Fig. 73. American Underfeed Stoker. 



of feeding the fuel to the retort and in the employment of " live " 
grates instead of dead plates on the sides of the retort. The coal is fed 
into the hopper and carried by an endless screw conveyor into the 
magazine or retort. Forced draft is used and the rate of draft and the 
speed of the conveyor are readily adjusted to suit the conditions of 
load. Underfeed stokers are very compact, occupy but little' space in 
front of the boiler, and are low in first cost. Careful manipulation 
is necessary to render them smokeless and efficient, since they are 
ordinarily installed without fire-tiled combustion chambers. 

93. Down-Draft Furnaces. — Fig. 74 shows the application of a 
Hawley down-draft furnace to a Heine water-tube boiler. In this 
furnace there are two separate grates, one above the other, the upper 



140 



STEAM POWER PLANT ENGINEERING 



one being formed of paral- 
lel water tubes connected 
with the water space of 
the boiler through the 
steel headers or drums A 
and D, in such a manner 
as to insure a positive 
circulation. Fuel is sup- 
plied to the upper grate, 
the lower one, formed of 
common bars, being fed 
by the half -consumed fuel 
falling from the upper 
grate. Air for combustion 
enters the upper fire door, 
which is kept open, and 
passes first through the 
bed of green fuel on the 
upper grate and then over 
the incandescent fuel on 
the lower grate. A strong 
draft is required, due to 
the relatively small upper- 
grate area and the corre- 
spondingly high rate of 
combustion. The down 
draft is very well adapted 
to the burning of paper, 
cardboard, excelsior, wood, 
and other rapidly burning 
refuse, as well as the vari- 
ous kinds of coal. Lump 
coal gives better results 
than the smaller sizes, as 
the latter are apt to fall 
through the upper grate 
before even partially con- 
sumed and when such is 
the case efficient results 
cannot be obtained. If 
carefully manipulated this furnace with fire-tiled tubes as illustrated 
in Fig. 74 gives high boiler efficiency and smokeless combustion, 




SMOKE PREVENTION, FURNACES, STOKERS 141 

but its overload capacity is limited. Without the fire tiling smokeless com- 
bustion is possible only at light loads. Hawley down-draft furnaces are in- 
stalled in a large number of tall office buildings in New York and Chicago 
and are giving excellent results. Down-draft furnaces are necessarily 
hand fired, since mechanical stokers are not easily adapted to them. 

94. Sprinkling Stokers. — In this system of stoking the fuel in finely 
divided form is distributed by sprinkling uniformly over the entire 
area of the grate. With the proper adjustment of air supply and feed 
the volatile gases are distilled off continuously before the grate is 
covered by the new coal and without materially lowering the tempera- 
ture of the incandescent fuel bed. Mechanically the operation involves 
considerable difficulty. One of the most successful American devices 
for this purpose is the " Little Giant " stoker. Coal of nut size or 
smaller is hand fed into a small hopper from which it gravitates on to a 
feed wheel driven by an engine or motor. A stream of coal is dis- 
charged into a cast-iron chute extending over the front part of the 
grate, from which it is blown into the furnace by a steam jet. The fine 
or powdered coal is burned in suspension and the heavier coal falls to 
the grates. A fire-brick combustion chamber is usually necessary for 
smokeless combustion, since the quantity of fuel burned may be too 
large to permit complete distillation of the green fuel before the suc- 
ceeding charges are delivered. A test on an English stoker of this type 
(Bennis stoker) , using compressed air for feeding the fuel, gave the unusu- 
ally high rate of combustion of 72 pounds of coal per square foot of grate 
surface per hour. The same test credited the boiler and grate with an 
efficiency of 84.9 per cent, which is probably the highest recorded 
efficiency of any boiler and furnace using coal as fuel. {Engineering 
Record, April 8, 1905, p. 404.) 

95. Dutch Ovens. — An independent furnace or Dutch oven in 
front of the boiler as illustrated in Fig. 75 provides one of the simplest 
methods of securing a large combustion chamber for the mingling of 
the air and combustible gases before delivering them to the boiler proper. 
Such a furnace produces very high temperatures when operating under . 
best conditions, and hence must be lined with fire brick of excellent 
quality. Although better than the ordinary setting the plain Dutch 
oven is too limited in length and capacity to prevent smoke from form- 
ing, except at very light loads. The velocity of the gases is usually 
too high to permit either a thorough mixture or complete oxidation 
before striking the boiler tubes. Steam jets placed at the sides of the 
setting and blowing across the fire are sometimes effective in mixing 
the air and combustible gases, but the best results are obtained by 
modifying the construction of the furnace to the extent of introducing 



142 



STEAM POWER PLANT ENGINEERING 



baffle walls which vary the direction of flow and by increasing the 
length of the path of the heated gases. The greater the length of the 
path and the greater the number of baffles the more thoroughly will the 
air and gases be mingled, but the intensity of draft will of course be de- 
creased in proportion. A compromise must therefore be made between 
required draft and length of path. The larger the extent of fire-tile sur- 
face the greater will be the regenerative effect, which is of particular im- 
portance in hand firing when the evolution of volatile gas is intermittent, 
but the first investment and cost of repairs and renewals are greater. 
There are little reliable data available pertaining to the relation between 
capacity of furnace and length of path of the heated gases for maximum 







Fig. 75. Plain Dutch Oven. 



efficiency. A modified Dutch oven is illustrated in Fig. 71. The exten- 
sion front is not necessary with some type of boilers, as will be seen 
from Figs. 63 and 74, in which a tile roof and baffles suitably arranged 
within the setting proper simulate the Dutch oven effect. (See " Cost 
of Maintenance of Dutch Oven Furnaces," A. Bement, Eng., U. S., 
July 1, 1907, p. 606.) 

96. Twin-Fire Furnace. — This arrangement, illustrated in Fig. 76 
in connection with a hand-fired return tubular boiler, is a double furnace 
formed by longitudinal arches extending between bridge wall and fire 
door. 

The furnaces are fed and manipulated alternately, the object being 
to have one furnace in a highly incandescent state while green fuel is 
fed into the other. Air is admitted both below and above the grate, 
and the volatile gases are supplied with the necessary oxygen for com- 



SMOKE PREVENTION, FURNACES, STOKERS 



143 



3dld Q33J 



J JO moia 



bustion before they come in contact with the comparatively cool boiler 
surface. 

The gases from both fur- 
naces first pass into a cham- 
ber formed by a single arch 
sprung across the entire inner 
setting from the side wall, a 
short retarding arch being 
placed between this interme- 
diate chamber and the rear 
of the setting. A special tile 
of high-grade refractory clay 
is used, the thickness varying 
from 4 to 6 inches, depending 
upon the size of furnace and 
the length of span. The fur- 
nace can readily be substi- 
tuted for the ordinary types 
in common use under any 
standard tubular or water- 
tube boiler and may be in- 
stalled either under the boiler, 
as indicated in the illustra- 
tion, or in an extension Dutch 
oven. This is an excellent 
furnace, and when properly 
manipulated gives smokeless 
and efficient combustion. 

96a. Chicago Settings for 
Hand-Fired Return Tubular 
Boilers. — Figs. 67a, 67b and 
67c show the general details 
of settings for return tubular 
boilers as recommended by 
the Chicago Department of 
Smoke Inspection. The set- 
ting illustrated in Fig. 67a is 
ordinarily installed where a 
strong draft is available and 
that shown in Fig. 67b or 67c 
where the draft conditions are not favorable. All three settings require 
careful manipulation for smokeless combustion as is the case with hand- 




144 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 



145 




146 



STEAM POWER PLANT ENGINEERING 




SMOKE PREVENTION, FURNACES, STOKERS 147 

fired furnaces in general. It has been the experience of the Department 
that most violations of the smoke ordinance are due primarily to insuffi- 
cient draft, the required rate of combustion being too high for the 
available air supply. The requirements outlined in paragraph 95 apply 
equally well to these settings. The following specifications refer to 
Figs. 67a, 67b and 67c, the items in the specifications corresponding 
to the letters in the illustrations. 

A. Doors should be of a type allowing the admission of excess air over 

the fire when so desired. If panels are cut in the fire doors for 
this purpose, the aggregate area of the openings should be not 
less than 4 square inches to each square foot of grate surface. 

B. Arches should be made of wedge brick or " bull heads " and not 

laid in two courses of 4j-inch brick. 

C. The bridge wall should be made of first grade fire brick above the 

grate line and with fire brick facing not less than 9 inches 
in thickness on the combustion chamber side. The top row 
should be a row-lock course. Provision should be made in 
the building of the bridge wall for lateral expansion. 

D. The combustion chamber floor should be paved with fire brick 

laid on edge. 

E. Fire brick lining below the arch skew-backs should be not less 

than 9 inches in thickness. Fire brick lining above the arch 
system and behind the deflection arch may be 4J inches in first 
grade fire brick, with headers every fifth row. 

F. Fire brick over firing door liners should be arched. This rule 

also applies to brick above the clean-out door openings. 

G. Facilities for taking up arch thrust should be provided in every 

case by suitable metal re-enforcements extending horizontally 
throughout the length of the arches. No air space should inter- 
vene between the metal re-enforcement and the skew-backs. 

H. Herringbone or Tupper grates or other similar types should not 
be selected where bituminous coal forms the major portion of 
the fuel. 

I. The back arch is preferably sprung from side to side rather than 
from back wall to rear boiler tube sheet. No metal should be 
exposed to direct heat of gases. 

J. Chimney heights of less than 75 feet above the grate line should 
not be permitted, and this height allowed only when the 
chimney is direct connected to the boiler uptake. In case of 
a breeching and detached chimney, add to the height of chim- 
ney computed by standard methods (never less than 75 feet) 
10 feet for every turn of the breeching and one foot for each 
foot in length of the breeching. 



148 



STEAM POWER PLANT ENGINEERING 



K. 



For boilers 48 inches or less in diameter, special provision for 
the examination of girth seams must sometimes be made. 
This is because of the fact that with small boilers there is not 
sufficient room between the arch and shell for purposes of 
inspection. 

UPTAKE 



^ , 







LONGITUDINAL SECTION 




SECTIONAL PLAN 

Fig. 77. Wooley Smokeless Furnace. 



L. In the event of arch failures, the boiler should be immediately 
taken out of service. This is to avoid failure of the boiler 
shell due to heat being applied upon a portion of the heating 
surface over which a mud deposit has formed. 



SMOKE PREVENTION, FURNACES, STOKERS 



149 



97. Wooley Smokeless Furnace. — Fig. 77 shows a longitudinal sec- 
tion and a sectional plan of a Wooley smokeless furnace applied to a 
B. & W. boiler. The main features of the furnace are a dividing wall 
in the fire box and a deflecting wall in the combustion chamber. The 
dividing wall permits of the alternate method of firing, whereby one 
side of the furnace is always in an incandescent state while the other 
side is being supplied with green fuel. If a mechanical stoker is used 
the wall in the fire box is omitted. The products of combustion are 
intended to be thoroughly mingled with the requisite amount of air by 
the deflecting walls before entering the regenerative or secondary com- 
bustion chamber. 




Fig. 78. Kent's Wing Wall Furnace. 



98. Kent's Wing- Wall Furnace. — Fig. 78 shows the application of 
Kent's wing-wall furnace to a water-tube boiler. The Dutch oven 
in front of the regular setting contains the grates. Wing walls F are 
placed as shown two or three feet to the rear of the bridge wall D, and 
fire-brick piers H behind the wing walls. 

In operation, fresh coal is spread alternately over each half of the 
grate. The dense smoky gases which rise from the green portion of 
the fire mingle in the narrow passage with the highly heated air which 
comes through the other side of the grate greatly in excess of that 
required to consume the partially burned coal there. The piers H act 



150 



STEAM POWER PLANT ENGINEERING 



as regenerative surfaces, absorbing heat from the fire when it is hottest 
and giving it out when it is coolest, that is, just after firing. 



E;£: 



I ttSSYflN 



AIR *22Z 

m 




mm 



Fig. 79. Burke's Smokeless Furnace, Front Section. 




BOILEK 



77777777Z777777 



''»'//////'/////,„„„„„»„„,, 



Fig. 



Burke's Smokeless Furnace, Side Section. 



Comparative tests of boilers with standard setting and with wing- wall 
furnaces have shown a much higher efficiency with the latter and with 
practically smokeless combustion. {Iron Trade Review, July 7, 1904, 
p. 76.) 



SMOKE PREVENTION, FURNACES, STOKERS 



151 



99. Burke's Smokeless Furnace. — Figs. 79 and 80 show sections 
through a Burke smokeless furnace as installed in a number of tall 
office buildings in Chicago. It amounts virtually to a Dutch oven 
equipped with shaking grates, and embodies an extension self-feeding 
coking oven of cast-iron section lined with fire brick and protected 
from overheating by air circulation through the sections. Natural 
draft is used, the fire doors being closed; but air is admitted above as 
well as below the fire. As this stoker is manipulated by hand, more or 
less attention is required of the operator in keeping the fire clean. 
Furnaces of this type at the power plant of the Majestic Theatre building, 
Chicago, 111., are giving excellent results. 

100. Admission of Air above Fire. — Smoke is often due to insuf- 
ficient air supply or imperfect mixing, especially when coal of a coking 
or clinkering nature is used 

which tends to seal up the air 

spaces in the grate. In these >H 

cases the admission of air above ^ 

the grate through openings in 

the bridge wall or passages in 

the side walls frequently gives 

satisfactory results. When 

natural draft is not sufficient, 

as is usually the case under 

heavy load, steam jets or forced draft may be employed. For a 

description of such devices see paragraphs 148 to 150. 

101. Cost of Stokers. — The following is the approximate cost of stokers 
suitable for a Babcock & Wilcox boiler of 350 horse power rated capacity 
with 45 square feet of grate surface ; height of chimney above grate, 1 75 feet ; 
coal burned, Illinois screenings. The cost of installation is not included. 

1. Chain grate and appurtenances $1,500.00 

2. Jones underfeed stoker 1,400.00 

3. Hawley down-draft furnace 1,350.00 

4. Burke smokeless furnace 1,000.00 

5. Roney stoker 1,300.00 

6. Murphy furnace and stoker 1,350.00 

7. Wilkinson stoker 1,200.00 

Parsons' Smokeless Furnace. See par. 149. 

Heinrich Smokeless Furnace. See par. 150. 
Steam jets. See par. 148. 
Hamler-Eddy Smoke Recorder. See par. 411a. 
Ringlemann Smoke Chart. See p. 765. 

Smoke Prevention: Bulletin No. 15, Univ. of 111., Vol. 
Geological Survey; Boiler Maker, May, 1909, Oct., 1909; 
Minn. Engr., Jan., 1910. 

Mechanical Stokers:Engr. U. S., Jan. 1, 1907, p. 83, Aug. 15, 1906, p. 540, July 2, 1906, 
p. 437;Cassier's Mag., Sept., 1906, p. 469; Power, Mar., 1906, p. 189, Aug., 1905, p. 487. 




Split Bridge Wall. 



431 ; Bulletin No. 334, U. S. 
Cassier's Mag., Feb., 1907; 



CHAPTER V. 

SUPERHEATED STEAM; SUPERHEATERS. 

102. General. — The steam engine fails to realize the efficiency of 
the ideal engine chiefly on account of cylinder condensation. The loss 
in heat due to this cause is seldom less than 10 per cent of the total 
supplied, and often as great as 40 per cent. 

If the steam is superheated before being admitted to the cylinder, 
condensation may be reduced or prevented entirely, as was recognized 
as early as sixty years ago, but the mechanical difficulties encountered 
prevented the practice until within the past few years. 

The principal advantages of superheated steam in connection with 
steam-engine work are: 

1. At high temperatures it behaves like a gas and is therefore in a 
far more stable condition than in the saturated form. Considerable 
heat may be abstracted without producing liquefaction, whereas the 
slightest absorption of heat from saturated steam results in condensa- 
tion. If superheat is high enough to supply not only the heat absorbed 
by the cylinder walls but also the heat equivalent of the work done 
during expansion, then the steam will be dry and saturated at release. 
This is the condition of maximum efficiency in a single cylinder. 
(Ripper, " Steam Engine Theory," p. 155.) Greater superheat than this 
will result in a loss of energy unless the steam is exhausted into 
another cylinder. To obtain dry steam at release the steam at cut off 
must be superheated 100 to 300 degrees F. above saturation tempera- 
ture, depending upon the initial condition of the steam and the number 
of expansions, a higher degree of superheat being required for earlier 
cut off. A superheat of 200 to 275 degrees F. at admission is necessary 
to insure dry steam at release in the average single-cylinder engine 
cutting off at one-fourth stroke, boiler pressure 100 pounds gauge. In 
most cases superheat is only carried so far as to reduce initial conden- 
sation, the steam becoming saturated at cut off, thus permitting efficient 
lubrication. There will be a reduction of approximately 1 per cent in 
cylinder condensation for every 7.5 to 10 degrees of superheat. In 
compound and triple-expansion engines the steam is ordinarily super- 
heated between each stage as well as before admission to the high- 
pressure cylinder. 

152 



SUPERHEATED STEAM; SUPERHEATERS 153 

2. A moderate amount of superheat produces a large increase in 
volume, the pressure remaining constant, and diminishes the weight of 
steam per stroke for a given amount of work. For example, the volume 
of 1 pound of saturated steam at 150 pounds pressure (gauge) is 2.75 
cubic feet, and its temperature is 365.8 degrees F. The total heat of one 
pound of this steam above the freezing point is 1193.5 B.T.U. By 
adding 110 B.T.U. in the form of superheat its temperature will be 
increased to 565.8 degrees F. (superheated 200 degrees F.) and its volume 
to approximately 3.5 cubic feet (specific heat taken as 0.55).* Thus 
an increase of 9.2 per cent in the heat effects an increase of 22 per cent 
in the volume, which means a corresponding reduction in the steam 
admitted to the engine per stroke. These figures are purely theoretical, 
as no allowances have been made for condensation of the saturated 
steam or for reduction in temperature of the superheated steam. 

3. Superheated steam has a much lower thermal conductivity 
than saturated steam, and therefore, less heat is absorbed per unit of 
time by the cylinder walls. 

General Discussion of Superheated Steam: Engr., Lond., Dec. 31, 1909; Eng., 
Lond., Sept. 13, 1901, Sept. 4, 1903, p. 237; Eng., U.S., Dec. 15, 1902, p. 821, 
Oct. 15, 1906, p. 687; Engr. Mag., Feb. 1903, p. 778, Sept., 1903, p. 897, Feb., 
1904, p. 757, June, 1904, p. 436, March, 1905, p. 943, Nov., 1905, p. 271, May, 
1906, p. 269; Eng. Rec, July 8, 1905, p. 28; June 30, 1906, p. 783, July 28, 1906, 
p. 86; Power, Aug., 1904, p. 463, Sept., 1904, p. 558, Oct., 1904, p. 762, Jan., 1905, 
p. 23, Feb., 18, 1908, Serial; Eng., Lond., Jan. 8, 1904, p. 42; West. Elec, Nov. 
14, 1903, p. 369; Proc. A.S.M.E., May 14, 1908. 

103. Economy of Superheat. — Many comparative tests of engines 
using saturated and superheated steam under varying conditions of 
pressure and temperature have been made during the past few years, 
showing in most cases a gain in favor of superheat due to the reduction 
in steam consumption, but in some cases the extra investment and cost 
of maintenance neutralize this gain, resulting in an actual loss when 
measured in dollars and cents per horse-power hour. 

As far as steam consumption per horse-power hour is concerned, 
superheating usually increases the economy five to fifteen per cent 

* The most satisfactory equation for determining the specific volume of super- 
heated steam is that given by Knoblauch, Linde, and Klebe (Peabody, " Steam and 
Entropy Tables," p. 22): 

pv = 0.5962 T - p(l+ 0.0014 p) / 150 > 3 Q Q > Q0Q _ .0833 V 

p = pressure, pounds per square inch absolute. 
T = absolute temperature of the steam, degrees F. 
v = specific volume of superheated steam, cubic feet. 



154 STEAM POWER PLANT ENGINEERING 

and in some instances as much as forty, the latter figure referring to 
the more wasteful types of engines. A fair estimate of the average 
reduction in steam consumption per horse-power hour with moderate 
superheating, that is, 100 to 125 degrees F., based on continuous opera- 
tion of existing plants, is: 

Per Cent. 

1. Slow running, full stroke, or throttling engines, including direct 

acting pumps 40 

2. Simple engines, non-condensing, with medium piston speed, includ- 

ing compound direct acting pumps 20 

3. Compound condensing Corliss engines 10 

4. Triple-expansion engines 6 

A prominent European builder of engines guarantees steam con- 
sumption with highly superheated steam as follows: 

Pounds per I.H.P. hour. 

Single-cylinder condensing engines 13.5 

Single-cylinder non-condensing engines^ 15.5 

Compound condensing engines 10 

Triple-expansion condensing engines 8.75 

In comparing the performances of engines using saturated and 
superheated steam it is advisable to base all results on the heat con- 
sumed per horse power rather than on the steam consumption, since 
the latter is apt to give a false idea of the relative economies. The 
real measure of economy is the cost of producing power, taking into 
consideration all charges, fixed and operating, and the next best is the 
coal consumption per I.H.P. hour, but as a means of comparing the 
engines only, the heat consumption per horse power per hour or per 
minute is very satisfactory. 

See paragraph 181 for the influence of superheat on the economy of 
reciprocating engines and paragraph 193 for the influence on steam 
turbines. 

Economy of Superheat : Eng. Mag., Dec, 1904, p. 757, April, 1905, Sept., 1903, 
p. 108; Trans. A.S.M.E., 22-899; Engr. Rec, July 8, 1905, p. 28; Power, Sept., 
1904, p. 558, Oct., 1904, p. 598, Jan., 1905, p. 23; Cassier's, Nov., 1903, p. 18. 

104. Limit of Superheat. — In this country steam temperatures 
exceeding 500 degrees F. are seldom employed, while in Europe few 
if any plants are installed without superheaters, and 600 degrees F. 
is a common temperature. 

Experience has shown that with engines of ordinary design, slide- 
valves and Corliss, the temperature at the throttle should not exceed 
500 degrees F. This corresponds to a superheat of 160 degrees F. with 






SUPERHEATED STEAM; SUPERHEATERS 155 

steam at 100 pounds gauge pressure, and 130 degrees F. at 150 pounds. 
This degree of superheat insures practically dry steam at cut off in the 
better grade of engines. Just how far superheating can be carried with 
a given engine of ordinary construction can be determined by experiment 
only, but a temperature of 500 degrees F. is probably an outside figure 
and 450 degrees F. a good average. Higher temperatures are apt to 
interfere with lubrication and sometimes cause warping of the valves. 
With temperatures below 450 degrees F. no difficulties are ordinarily 
met with. Metallic packing has been found to give the best results for 
both piston rods and valve stem. 

It is generally assumed that a greater quantity of oil is required 
for lubricating valves and cylinders in connection with superheated 
steam, but experience seems to show that such is not the case. (Proc. 
A.S.M.E., May 14, 1908.) Forced-feed lubricators are the most satis- 
factory for superheated steam engines, since they insure a positive 
and copious flow of oil directly to the valves or other parts requiring it.* 

With highly superheated steam involving temperatures of 600 
degrees F. or more the poppet-valve type of engine is ordinarily 
employed, though balanced piston valves are not uncommon. The 
poppet valve is not distorted by heat and requires no lubrication. In 
Europe these engines have been brought to a high state of efficiency, 
but have not been generally adopted in this country owing, no doubt, 
to the higher cost. 

105. Specific Heat of Superheated Steam.f — The total heat of super- 
heated steam is given as „■ , , ~ . n ~y. 

ti = A -r Cpt, (li) 

in which 

X = B.T.U. in one pound of saturated steam above 32 degrees F. 

C p = mean specific heat of the superheated steam at constant pressure. 
t = degree of superheat, degrees F. 

Regnault determined the mean specific heat at atmospheric pressure 
to be 0.48 between 127 degrees and 226 degrees C. of superheating, and 
until recently this has been assumed to apply to all pressures and tem- 
peratures. As early as 1876 Hirn concluded from experiments made 
with a throttling calorimeter that the specific heat of saturated steam 
increased with the pressures and decreased at any given pressure if the 
steam became superheated. Since then numerous investigators have 
promulgated theories pertaining to this subject which have been far 
from harmonious and none has been universally accepted. Some 
experiments appear to show that specific heat is independent of pressure 

* Effect of Superheated Steam on Cylinder Oils. Mech. Engr., Lond., July 31, 
1908, p. 115. 

f See paragraph 113a. 



156 



STEAM POWER PLANT ENGINEERING 



and degree of superheat, while others indicate an increasing value as 
the pressure and degree of heat increase. Still others corroborate 
Hirn's theory. 



0.60 



0.55 



0.50 



0.46 













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roo 



Fig. 82. Specific Heat of Superheated Steam, Knoblauch and Jakob. 



-23A25. 



U 16 



— "77 P! 



IV 






40 



GO 



140 



160 



80 100 120 

Pressure-Lb. Abs. 

Fig. 83. Specific Heat of Superheated Steam, A. R. Dodge. 



180 200 



The maximum figure ranges as high as 0.8 and the minimum 0.48 for 
a given pressure and degree of superheat. 

The curves in Fig. 82 are based upon the experiments of Knoblauch 
and Jakob ("Mitteilungen uber Forschungsarbeiten," etc., Heft 36, 
p. 109, and Stevens' Indicator, October, 1905); those in Fig. 83 upon the 



SUPERHEATED STEAM; SUPERHEATERS 



157 



experiments of A. R. Dodge (Trans. A.S.M.E., 1907); those in Fig. 84 
are plotted from tests of Burgoon, Carpenter, and Thomas (Trans. 
A.S.M.E, 1907); and those in Fig. 85 are based upon the investigation 
of Professor Thomas (Trans. A.S.M.E., December, 1907). These curves 







\ 














Specific Heat of Steam 














*70 

.68 

.66 

^64 

■§ -62 

W.60 

§ .58 

t» .56 

.54 

52 






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Curves in Black :-C.E. Burgoon 
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40 



80 



400 



120 160 200 240 280 320 3( 

Degrees, F. of Superheat 
Fig. 84. Specific Heat of Superheated Steam, C. E. Burgoon. 

differ both in theory and in value of c p , but until further experiments 
prove otherwise the values in Fig. 82 may be accepted as sufficiently 
accurate for all practical purposes. The values given by Knoblauch 
and Jakob have been accepted by authorities as the most reliable. 
Table 17a is based upon their results. Table 17b has been calculated 
by means of Linde's equation. (See footnote, page 153.) 

TABLE 17. 

VALUE OF c p AT ATMOSPHERIC PRESSURE BY VARIOUS AUTHORITIES. 
Superheated Steam Cooled by Water- Jacketed Calorimeter. 





Publication and Date. 


Temp. 
Deg. F. 


c p at 

Atmos. 

Pres. 


Variation of c p with 




Increasing 
Pressure. 


Increasing 
Temp. 


Regnault 


Ann. de Chimie 
et de Physique, 
Tome 23. 

Sibley Journal, 

5-1904. 
Trans. A.S.M.E. . . . 


Varied 

Varied 
Varied 


0.4805 

0.4844 
0.48 


None 

Increases 
Increases 


None 
None 


Carpenter (Jones) 
Dodge 







158 



STEAM POWER PLANT ENGINEERING 



TABLE 17 — Continued. 
Throttling Calorimeter. Saturated Steam Expanded to Lower Pressure. 



Author. 


Publication and Date. 


Temp. 
Deg. F. 


c p at 

Atmos. 

Pres. 


Variation of c p with 




Increasing 
Pressure. 


Increasing 
Temp. 


Grindley 


Phil. Trans., Vol. 
194. 


239 


0.4317 


Increases 

Increases 
Increases 

Increases 

Increases 

Increases 

Increases 


Increases 


Hirn 


Griessmann 

Peake 


Zeit. V.D.Ing.,52, 

1903. 
Proc. Royal Soc. 

A-509, 1905. 
Sibley Journal, 

May, 1904. 
Sibley Journal, 

May, 1904. 
Sibley Journal, 

May, 1904. 


269 
Varied 
Varied 
Varied 
Varied 


0.506 

0.43 

0.463 

0.4825 

0.48 


Increases 

Increases 

None 

None 

None 


Carpenter (Stew- 
art and Marble).. 

Carpenter (Hoxie 
and Wood). 

Carpenter 
(Sickles). 



Superheating Steam Electrically. 



Peake 



Carpenter (Berry) 
Carpenter 
(Thomas). 



Lorenz 

Knoblauch and 
Jakob. 



Proc. Royal Society 
A509, 1903. 



Trans. A.S.M.E., 
Eng. Mag., March 
1907. 

Z.V.D.I.,No. 20... 
Engineering, L, 
Feb. 22, 1907. 



Varied 


0.46 


Varied 


0.48 


212 


0.49 


402 


0.487 


212 


0.445 


700 


0.49 



None 
Increases 



Increases 
Increases 



Decreases 



None 
None 



Decreases 

Decreases 
Decreases 

then 

increases 



From Combustion of Explosive Gases. 



Mallard and 

LeChatelier. 
Sarran and Vieille 
Langen 



Zeit. V. D. Ing. 
Tome 48. 

...Do 

...Do 



212 

212 
212 



0.46 

0.464 
0.463 



None 

None 
None 



Increases 

Increases 
Increases 



From Calculation. 



Reeve 


Wor. Poly. Journal 


215 
265 
236 


0.39 

0.4895 

0.38 

0.568 

0.468 

0.36 

0.4805 

0.513 

0.493 

0.479 


Increases 
Increases 


Increases 


Hirn. 


Decreases 




London Engineer . . 




Zeuner 






Weyrauch 

Perry 


Zeit. V. D. Ing., 
Tome 48. 

Steam Engine 

Thermodynamics . . 

Rose Technic, 1905 

Publication by au- 
thors, Berlin, 1905 


212 

212 
Varied 
284.4 
212 
356 


Decreases 

Increases 
None 
Increases 
Increases 


Increases 
None 


Roentgen 


None 






Knoblauch, Linde, 


Decreases 









SUPERHEATED STEAM; SUPERHEATERS 



159 



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160 



STEAM POWER PLANT ENGINEERING 



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162 



STEAM POWER PLANT ENGINEERING 



Just how much these different values of the specific heat affect the 
calculated performance of an engine may be illustrated by the following 
case: 

An engine uses 8.5 pounds of steam per I.H.P. hour, initial pressure 
120 pounds absolute; back pressure 0.94 pound absolute; superheat 
at throttle 300 degrees F. Assume that the " heat of the liquid " 
in the exhaust steam is returned to the boiler and that the specific 
heat is 0.48 in one case and 0.8 in the other, these being the extreme 
values given by different experimenters for the given conditions. 



Specific heat 

B.T.U. in 1 pound of saturated steam above 

ideal hot- well temperature 

B.T.U. in superheat 

B.T.U. in 1 pound of superheated steam . . 
B.T.U. per I.H.P. per minute 



Case I. 



Case II. 



1118.0 
144 


0.48 
162. 1 


1118.0 
240 


0.8 


1262.0 


1358.0 


192.4 



On the basis of the specific heat of 0.48 the heat consumption, 162.1 
B.T.U. per I.H.P. per minute, is exceptional and corresponds to an 
equivalent saturated steam consumption of 9.6 pounds per I.H.P. 
hour, whereas the performance of 192.4 B.T.U. per I.H.P. per minute 
based upon the specific heat of 0.8 has been excelled by a number of 
actual engines using saturated steam. 

106. Superheaters. — The installation of a superheater is equivalent 
to an increase in boiler capacity. The superheater may be independ- 
ently fired or can be arranged in connection with the boiler proper. 
The heating surface is usually of wrought iron, mild steel, cast iron, or 
cast steel. Engineers are not agreed as to which arrangement or which 
material gives the most economical returns. The recent symposium 
on superheated steam conducted by the research committee of the 
American Society of Mechanical Engineers (Trans. A.S.M.E., 1907) 
clearly indicated this lack of agreement. The requirements for a suc- 
cessful superheater are: 

1. Security in operation, or minimum danger of overheating. 

2. Economical use of heat applied. 

3. No exposure of joints to the fire. 

4. Provision for free expansion. 

5. Disposition such that it may be cut out or repaired without 
interfering with the operation of the plant. 

6. Ease of application to existing plants. 



SUPERHEATED STEAM; SUPERHEATERS 



163 



Nearly all superheaters depend upon carrying steam at a high velocity 
through small tubes in the form of return bends or coils and arranged 
to be heated by the hot gases in the boiler furnace or from some other 
source. 

The independently fired superheater has the following advantages: 

1. The degree of superheat may be varied independently of the per- 
formance of the boiler. 

2. It can be placed at any desirable point. 

3. Repairs are readily made without shutting down boilers. 

Some of the disadvantages are: 

1. It requires separate firing and extra attention. 

2. Saturated steam can only be furnished in case of a breakdown to 
the superheater. 

3. Extra piping is required. 

4. Extra space required. 

Standard practice in this country advocates that the superheater be 
contained within the boiler setting. Of two hundred recent installa- 
tions, one hundred and eighty, or ninety per cent, were of this type. 




Fig. 86. Babcock and Wilcox Superheater. 

107. Babcock & Wilcox Superheater. — Fig. 86 shows the applica- 
tion of superheating coils to a Babcock & Wilcox boiler, illustrating 
the indirectly fired type. The wrought-iron tubes are bent into U shape, 
the ends being connected into manifolds, the upper one receiving the 
saturated steam from the boiler and the lower one the superheated 



164 



STEAM POWER PLANT ENGINEERING 



steam after it has traversed the superheater tubes. A small pipe con- 
nects the lower manifold with the water space of the boiler by means of 
which the superheater may be cut out if desired, or flooded when starting 
up. Any steam formed in the superheater tubes is returned into the 
boiler drum through the collecting pipe, which, when the superheater 
is at work, conveys saturated steam into the upper manifold. When 
steam pressure has been attained the superheater is thrown into action 
by draining the water away from the manifolds and opening the super- 
heater stop valve. The tubes are free at one end and the manifolds are 
not rigidly connected with each other, thus avoiding expansion strains. 
With the proportion of superheating surface to boiler surface ordinarily 
adopted the steam is superheated from 100 to 150 degrees F, 



STEAM PIPE 



FEED PIPE 



SAFETY VALVE 




Fig. 87. Stirling Superheater. 

108. Stirling Superheater. — This superheater consists of two drums, 
Fig. 88, connected by seamless drawn tubes two inches in diameter. It 
may take the place of the middle bank of tubes in the Stirling boiler as 
shown in Fig. 87, or be installed in the final pass of the gases in the back 






SUPERHEATED STEAM; SUPERHEATERS 



165 



of the boiler. The drums around the tubes are protected from intense 
heat by asbestos cement. A pipe connecting the front drum of the 
boiler with the lower drum of the superheater permits the coils to be 
flooded in starting up or when the superheater is not needed. In this 




Fig. 88. Arrangement of Tubes ; Stirling Superheater. 



case the superheater acts as additional boiler-heating surface. The 
upper drum is divided into three and the lower into two compartments. 
The tubes are arranged with alternately wide and narrow spacing, so 
that any tube may be removed without disturbing the rest. The flow 
of steam is indicated by arrows. 

109. Foster Superheater. — Fig. 89 shows the application of a 
Foster superheater to a Babcock & Wilcox boiler. This device consists 
of cast-iron headers joined by a bank of straight parallel seamless 
drawn-steel tubes, each tube being encased in a series of annular flanges 
placed close to each other and forming an external cast-iron covering 
of large surface. The tubes are double, the inner tube serving to form 
a thin annular space through which the steam passes as indicated. 
Caps are provided at the end of each element for inspection and cleaning 
purposes. Foster superheaters are more costly than plain-tube super- 
heaters, but are longer lived and offer a much larger heating surface in 
proportion to the space occupied. 

Fig. 91 shows a Foster superheater arranged for independent firing. 



166 



STEAM POWER PLANT ENGINEERING 



The " Schwoerer " superheater, which is somewhat similar in external 
appearance to the Foster, differs from it considerably in detail, the 
heating surface being made up of suitable lengths of cast-iron pipe 
ribbed outside circumferentially and inside longitudinally. The ends 
of the pipes are flanged and connected by cast-iron U-bends. The 
intention is to provide ample heating surface internally and externally, 
with a compact apparatus. 




Fig. 89. Foster Superheater in Babcock & Wilcox Boiler. 



110. Independently Fired Superheaters. — The Schmidt superheater, 
Fig. 90, consists of two nests of coils, A and D, of equal size and dimen- 
sions, connected to cast-iron headers and 7. Saturated steam enters 
the first nest of coils through C and passes into header 0. From 
the steam, which is now dried and partly superheated, flows through 
the cast-iron pipe E to header I, and thence through the second nest 
of coils into header adjoining 0, and through pipe R to the engine. 
In chamber D the steam and gases flow on the counter-current and in 
chamber A on the concurrent principle. This combination permits of 
a low flue temperature and high steam temperature without subjecting 
the tubes to an excess of heat as would be the case if the steam left the 
coils A at header I, where the furnace gases are the hottest. A steam 



SUPERHEATED STEAM; SUPERHEATERS 



167 




168 



STEAM POWER PLANT ENGINEERING 




SUPERHEATED STEAM; SUPERHEATERS 



169 






temperature of 750 degrees F. and a flue temperature of 450 degrees F. 
are easily maintained with this apparatus. A mercury pyrometer T 
is fitted where the superheated steam enters the discharge pipe R. 
A thermometer cup L permits of checking the pyrometer by means of 
a nitrogen-filled thermometer. Each coil can be taken out separately 
and a new one put in without removing the others or dismantling the 
plant. Water produced by condensation while the superheater is 
inoperative collects in the bottom header N and escapes through a 
drain cock. If the steam supply should be suddenly shut off, the air 
door P is opened automatically by weight K. As soon as steam 
begins to flow it raises the weight through the opening of valve C and 
the door closes. The Schmidt superheater when arranged in the flue 
has practically the same construction as the independently fired. 



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FEED WATER 




ECONOMIZER 



Fig. 92. Schmidt System of Combined Superheater, Feed Water Heater and Economizer. 



Fig. 92 shows a combination of Schmidt superheater, economizer, 
and feed-water heater which finds much favor with engineers on the 
Continent. 



170 STEAM POWER PLANT ENGINEERING 

111. Materials used in Construction of Superheaters. — Most super- 
heaters are constructed either of wrought iron, mild steel, cast iron, or 
cast steel, the latter material having the advantage of not being dam- 
aged by any temperature to which it is likely to be subjected, which 
does away with the necessity of damper mechanisms and simplifies the 
installation. On the other hand, cast-metal superheaters are usually 
ribbed after the fashion of an air-cooled gas engine, and are, therefore, 
very heavy and thick walled, necessitating a higher temperature for 
the same useful effect than in the case of the wrought-iron construction, 
but have the advantage of minimizing fluctuation of steam temperature 
which would otherwise be caused by a wide variation in temperature of 
furnace. One of the most successful cast-metal heaters is of European 
design and is constructed of a special alloy known as " Schwoerer " 
iron. Table 18 gives the yearly cost of repairs to piping and necessary 
brickwork for a number of installations equipped with cast-metal super- 
heaters of the " Schwoerer " type. 

Wrought iron and mild steel offer the advantage of lightness, ease of 
construction, and low first cost, but cannot be exposed to very high 
temperatures without injury, and consequently provision must be made 
for diverting the direction of the heated gases or for flooding the coils 
while the boiler is being warmed before steam is generated. 

Neither cast iron nor steel loses in tensile strength when subjected for 
a very short time to the temperature of superheated steam, but, on the 
contrary, may be stronger. Tests made by Professor Lanza (" Applied 
Mechanics," p. 489) showed that the tensile strength of steel dimin- 
ished from degrees F. to about 300 degrees F. and then increased, 
reaching a maximum between 500 and 650 degrees F. Cast iron and 
steel maintained their strength, with a tendency to increase, up to 900 
degrees F. beyond which the strength is diminished. 

Ordinary cast-iron valves and fittings have shown permanent increase 
in dimensions under high superheat and in numerous instances have failed 
altogether, but sufficient data are not available to prove conclusively the 
unreliability of cast iron if the iron mixture is properly compounded 
and the necessary provision is made for expansion and contraction.* 

112. Extent of Superheating Surface. — The required extent of super- 
heating surface for any proposed installation depends upon (1) the 
degree of superheat to be maintained; (2) the velocity of the steam 
through the superheater; (3) the character of the superheater; (4) the 
weight of the steam to be superheated; (5) the moisture in the wet steam; 
(6) the temperature of the gases entering and leaving the superheater; 
and (7) the conductivity of the material. 

* See Symposium on the "Effect of Superheated Steam on Cast Iron and Steel," 
Jour. Am. Soc. Mech. Engrs., Dec., 1909. 

. 



SUPERHEATED STEAM; SUPERHEATERS 



171 



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172 STEAM POWER PLANT ENGINEERING 

Since the heat absorbed by the steam in the superheater is equal to 
that given up by the products of combustion, neglecting radiation, this 
relationship may be expressed 

SUd = Wc ft - * 2 ), (23) 

in which 

S = square feet of superheating surface per boiler horse power. 
U = coefficient of heat transmission, B.T.U. per square foot per 

hour per degree difference in temperature. 
d = mean temperature difference between the steam and heated 

gases, degrees F. 
W = weight of gases passing through the superheater per boiler 

horse power hour. 
c = specific heat of the gases. 

t t = temperature of the gases entering superheater, degrees F. 
t 2 = temperature of the gases leaving superheater, degrees F. 

Transposing equation (23), 

8 = Wc (*'- '»> , (24) 

Ud 

On account of the great variation in the values given for U, and 
the difficulty of determining d, t Xi and t 2 for different types of super- 
heaters, equation (24) is hardly applicable in practice. 

An empirical formula for determining the extent of superheating 
surface in connection with indirect superheaters which appears to con- 
form with practice is given by J. E. Bell (Trans. A.S.M.E., May, 1907) : 

10 S (25) 



2{T - t) - S 
in which 

x = square feet superheating surface per boiler horse power. 
S = superheat, degrees F. 

T = temperature of the products of combustion where the super- 
heater is located. 
t = temperature of the saturated steam. 

The value of T may be found from the equation 
1 



(T - t) ' 16 
in which 



= 0.172 H + 0.294, (26) 



H = the per cent of boiler-heating surface between the point at 

which the temperature is T and the furnace. 
t as in (25). 






SUPERHEATED STEAM; SUPERHEATERS 



173 




83SVO as U3AO oassvd aovjuns onuvsh uxlvm so xn30 msa 



174 STEAM POWER PLANT ENGINEERING 

Equation (26) is based upon the assumption that the heat trans- 
ferred from the gases to the water is directly proportional to the differ- 
ence in temperature; that the furnace temperature is 2,500 degrees F.; 
flue temperature 500 degrees F.; steam pressure 175 pounds per square 
inch gauge; one boiler horse power is equivalent to 10 square feet of 
water-heating surface. 

Example: What extent of heating surface is necessary to superheat 
saturated steam at 175 pounds gauge pressure, 200 degrees F., if the 
superheater is placed in the boiler setting where the gases have already 
traversed 40 per cent of the water-heating surface? 

Substitute H = 0.4 and t = 378 in equation (26), 

= 0.172 X 0.4 + 0.294 



(T - 378)°' 

T = 950. 

Substitute T = 950 and S = 200 in equation (25), 

= 10 X 200 

" 2 (950 - 378) - 200 

= 2.12 square feet. 

The curve in Fig. 92a was plotted from equation (26), and gives a 
ready means of determining T and of observing the law governing 
heat absorption by the boiler between furnace and breeching. The 
abscissas represent the temperatures of the hot gases at different points 
in their path between furnace and breeching. The ordinates represent 
(1) the per cent of boiler-heating surface passed over by the hot gases, 
and (2) the per cent of the total heat generated which is absorbed by 
this heating surface. 

In the use of equation (26) the probability of error is greatest when 
considering a point near the furnace, since large quantities of heat are 
transmitted to the tubes by radiation from the fuel bed which are not 
taken account of. For most practicable purposes the assumption is 
sufficiently accurate. 

For the application of the curve in Fig. 92a to the design of direct 
and indirect superheaters for various degrees of superheat, see " Stir- 
ling," published by the Stirling Boiler Company, pp. 92-96. 

113. Performance of Superheaters. — Published tests of both directly 
and indirectly fired superheaters cover such a wide range of conditions 
of installation and operation that general conclusions cannot be drawn, 
but it may be of interest to note briefly the performances in a few 
specific cases. 



SUPERHEATED STEAM; SUPERHEATERS 



175 



The curves in Figs. 93, 94, and 95 are plotted from tests of a Babcock 
& Wilcox boiler, with 5000 square feet of water-heating surface, 
equipped with superheating coils of 1000 square feet area, as illustrated 
in Fig. 62. The furnace with ordi- 
nary short ignition arch was pro- 
vided with chain grate of 75 
square feet area. 

Fig. 93 shows the relation be- 
tween degrees of superheating and 
total horse power of boiler and 
superheater. 

Fig. 94 shows the relation be- 
tween the horse power produced Horse Power Produced in Boiler 

Fig. 94. Percentage of Horse-Power Pro- 
duced in the Superheater of that Devel- 
oped in the Boiler. 



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to Total Horse-Power Developed. 



Fig. 95. Relation of Degree of Superheat 
to Horse-Power of Superheater. 



in the boiler and the percentage of boiler horse power produced in the 
superheater. 

Fig. 95 shows the relation between the degree of superheat obtained 
and the horse power developed in the superheater. 



176 



STEAM POWER PLANT ENGINEERING 



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SUPERHEATED STEAM; SUPERHEATERS 



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178 



STEAM POWER PLANT ENGINEERING 



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SUPERHEATED STEAM; SUPERHEATERS 



179 



Tables 19 to 21 are taken from the report of Otto Berner (" Zeit. d. 
Ver. Deut. Eng." and reprinted in Power August, 1904). 

Table 19 compares the heat efficiency of a steam plant equipped with 
directly and with separately fired superheaters, the former showing a 
much higher efficiency. 

Table 20 compares different boilers with and without flue super- 
heaters, showing the effect upon the temperature of the flue gases. 
The gain in heat efficiency of the entire plant due to the use of the super- 
heater is decisive in each case. 

TABLE 22. - 

(Engineer, U. S., May 1, 1904.) 



Time of start. 
Time of finish . 



Hours run 

Average steam pressure 

Average water pressure, triple expansion 

head in feet 

Average water pressure, compound, head 

in feet 

Average vacuum of suction for triple and 

compound, inches of mercury 

Total head on triple, feet of water 

Total head on compound, feet of water . . . 

Total double strokes, triple 

Total double strokes, compound 

Gallons pumped from piston displacement, 

total, triple 

Gallons pumped from piston displacement, 

total, compound 

Gallons pumped from piston displacement, 

total, triple combined 

Gallons, total, pumped as measured by weir 

Per cent slip 

Foot pounds, weir 

Total coal consumed 

Per cent refuse 

Total refuse 

Total feed water 

Duty per 100 pounds coal 

Duty per 1,000 pounds steam 



With Superheater. 


Without Superheater. 


12 noon, Feb. 8 


11 a.m., Feb. 11 


12 noon, Feb. 9 


11 a.m., Feb. 12 


24 


24 


79.3 1b. 


79.4 1b. 


0.99 


1.05 


7.10 


7.10 


22.90 


23.21 


29.05 


29.46 


33.04 


33.39 


30,557 


34,114 


35,395 


32,158 


2,854,023 


3,186,247 


2,930,706 


2,662,682 


5,784,720 


5,848,930 


4,492,680 


4,549,480 


22.3 


22.2 


1,163,815,819 


1,184,983,596 


5,015 lb. 


6,410 lb. 


23.7 


18.7 


1,188 


1,203 


38,399 


50,960 


23,206,696 


18,486,483 


30,308,498 


23,253,213 



Per cent increase of work per 100 pounds coal 25.5 

Per cent increase of work per 1,000 pounds steam 30.2 

Per cent saving in coal per foot pound work 20 . 2 

Per cent saving in feed water per foot pound work 23 . 2 

Average temperature steam leaving superheater 527 . 4 deg. F. 

Average temperature steam entering superheater 320. 1 deg. F. 

Average degree superheat 207 . 3 deg. F. 



180 STEAM POWER PLANT ENGINEERING 

Table 21 shows the gain in heat efficiency due to the use of super- 
heaters in a number of plants equipped with fire-tube boilers. 

Table 22 gives the results of tests on one of the return tubular boilers 
at the Spring Creek Pumping Station of the Brooklyn Waterworks 
(leb. 9, 1904) with and without a superheater. The superheater, 
of the Foster type, was installed between the rear wall of the setting 
and the tube sheet. 

113a. Properties of Superheated Steam. — The following equations 
derived by Prof. Goodenough of the University of Illinois and based 
upon the experiments of Knoblauch and Jakob, give a comparatively 
simple method of determining the various properties of superheated 
steam if steam and entropy tables are not available. The results as 
obtained from these equations agree substantially with Marks' and 
Davis' Steam Tables and the 1909 Edition of Peabody's Steam Tables. 

T = absolute temperature of the superheated steam, deg. F. 
p = absolute steam pressure, lbs. per sq. in. 
X = total heat, B.T.U. per pound. 
u = intrinsic energy, ft. -lbs. 
n = entropy. 
C p = true specific heat. 
X=T (0.372 + 0.00005 T)-p(l + 0.00035) -^ + 882.4, 

in which log C = 9.42383. 
u = T (202.44 + 0.0389 T) - j^. (1 + 0.00025 p) + 686242, 

in which log C = 12.20551. 
n = 0.85657 log T + 0.0001 T - 0.25392 log p 

- p (1 + 0.00035 p) £ M - 0.4300, 
in which log C = 9.31469. 
C p = 0.372 + 0.0001 T - p (1 + 0.00035 p) ^- 5 , 
in which log C = 9.96790. 

The mean specific heat may be obtained by subtracting the total heat 
of the saturated steam from that of the superheated steam and dividing 
the difference by the degree of superheat. 

The specific volume may be determined from Linde's equation as 
stated at the bottom of page 153. 






CHAPTER VI. 

COAL AND ASH-HANDLING APPARATUS. 

114. General. — The cost of coal and its delivery into the furnace 
are usually the largest items in the operating charges, hence large 
central stations are located, when practicable, adjacent to a railway 
line or water front, to minimize the cost of handling coal and ashes. 
Isolated stations in the business districts of large cities are usually 
unfavorably situated, so that the cost of handling coal and ashes is a large 
percentage of the total fuel cost. In large stations the amount of fuel 
and ash handled frequently warrants the expense of elaborate conveyor 
systems which would not be justified in smaller plants. In whatever 
way coal is supplied provision should be made for storing a quantity 
sufficient to operate the plant for some time in case the supply is inter- 
rupted, thereby guarding against an enforced shut-down. 

If adjacent to a railway line, a side track must be provided for switch- 
ing the cars. As bottom-dumping cars cannot be depended upon, 
provision should be made for unloading by hand. If coal is delivered 
by water, clam-shell drop buckets are ordinarily used for unloading 
the barges. If the power house is located at some distance from the 
railroad or water the coal is generally hauled by teams in two to five- 
ton loads. 

115. Coal Storage. — In small stations the storage bins or coal 
bunkers may usually be located within the building, but in larger plants 
the quantity of coal consumed daily is frequently such that an immense 
space would be required to furnish storage capacity for even a short 
period of time. For example, one of the large central stations in 
Chicago burns an average of 30 tons of Illinois screenings per hour 
throughout the year. Allowing 45 cubic feet to the ton this would 
necessitate a space of 45 x 30 x 24 = 34,800 cubic feet to store coal for 
one day's operation. A ten-days run would require a coal pile 50 feet 
wide, 30 feet high, and 232 feet long. It is a good plan, if the location 
and character of the plant permit, to carry four or five days' supply 
within the plant and provide a separate building for the coal reserve. 
Such provision is made in the power plant of the New York Edison 
Company, which has a storage capacity of 150,000 tons in addition to 
that of the overhead bunkers. 

181 



182 STEAM POWER PLANT ENGINEERING 

Exposed coal piles are objectionable, because of freezing in winter, the 
crust sometimes freezing so hard as to necessitate the use of dynamite 
to break it; moreover, a slow depreciation in heat value takes place, 
especially with bituminous coal. This depreciation is more rapid in 
warm weather and in the tropics. Stored coal is oftentimes subject to 
spontaneous combustion, particularly when there is a large content of 
iron pyrites. 

Coal bunkers or hoppers are ordinarily placed on the same level with 
boiler-room floor or above the boiler setting. The former is the cheaper 
as far as first cost is concerned, but necessitates additional handling of 
the fuel before it can be fed to the stokers. In the overhead system 
the coal gravitates to the stoker through down spouts. Overhead 
bunkers are usually found where real estate is costly. They are gener- 
ally constructed of steel plates lined with concrete or of reenforced 
concrete. The bottoms slope at an angle of 35 to 45 degrees and empty 
into the coal chutes or down spouts. Fig. 99 shows the general appear- 
ance of a single overhead bunker and Fig. 441 that of a double bunker. 
In some bunkers the floors are made with very slight slopes, but it is not 
advisable to use a slope less than the angle of repose of the coal, as it may 
be necessary to shovel the coal over the spouts. Convenience in framing 
makes the 45-degree slope the more desirable. Separate bunkers for 
each boiler are preferred to continuous bunkers, since fire in the coal is 
more readily prevented from spreading. In the new power house of 
Swift & Co., Chicago, 111., the bunkers are of circular cross section 
instead of rectangular as is the usual practice. The capacity of the 
cylindrical hopper is considerably less than that of a rectangular 
hopper of the same width, but is much cheaper to construct. 

Ash bins are invariably lined with concrete or brickwork, since the 
corrosive action of the ashes would soon destroy the bare iron, and are 
usually located alongside the coal hopper, as in Figs. 97 and 99, so that 
they may be discharged by gravity. The angle of repose of most ashes 
is approximately 40 degrees, but the 45-degree angle is preferred on 
account of convenience in construction. 

Coal Storage: Power, April, 1907, p. 217, Aug., 1899, p. 3, Nov., 1904, p. 651; 
Eng. Rec, Sept. 23, 1905, p. 534, June 1, 1902, p. 532, July 4, 1903, p. 4; Eng. 
News, July 11, 1907, June 5, 1902, p. 463, April 3, 1903, p. 272; West. Elec., Oct. 
28, 1905, p. 335; Trans. A.S.M.E., 23-473. 

Coal Storage under Water: Eng. News, Dec. 24, 1908; Eng. Min. Jour., Dec. 1, 1904, 
p. 975; Engineering, Sept. 4, 1903, p. 863. 

Calorific Value of Weathered Coals: Trans. A.S.M.E., 20-333; Bulletin No. 17, 
Univ. of 111., Aug. 26, 1907. 

Design of Coal and Ash Bins : Eng. News, July 21, 1904, p. 62; Eng. Rec, Sept. 
1, 1900, p. 201; Power, Nov., 1899, p. 14, Nov. 1904, p. 651; Elec. Age, March, 
1907, p. 141. 






COAL AND ASH-HANDLING APPARATUS 183 

116. Coal Conveyors. — Coal is carried to the stokers in a variety of 
ways, depending upon the location of the plant, the type of stokers, 
and the personal tastes of the builder. Of the various methods the 
following are the most common : 

1. Hand shoveling from coal pile to furnace. 

2. Wheelbarrow or hand car and shovel. 

3. Bucket conveyor. 

4. Belt conveyors. 

5. Hoist and hand cars. 

6. Hoist and automatic cable cars. 

7. Combinations of the above. 

Coal-Handling Plants for Power Houses : Am. Elec, June, 1900, p. 266, Oct., 
1901, p. 486; Cassier's, April, 1905, p. 480; Elec. World., Dec, 1901, p. 463; Engr., 
U.S., July 1, 1904, p. 461; Jan. 1, 1905, p. 4; Eng. Rec, April 5, 1902, p. 322. 

Ash Handling : Elec. World., Oct. 5, 1901, p. 569; Eng. News, Oct. 19, 1905, 
p. 403; Eng. Rec, May 10," 1902, p. 435, Jan. 17, 1903, p. 85, Feb. 7, 1903, p. 153, 
Oct. 28, 1905, p. 482, Oct. 7, 1905, p. 396, Dec. 9, 1905, p. 655; Power, Oct., 
1904, p. 507, July, 1904, p. 422; St. Ry. Jour., Jan. 5, 1901, p. 11. 

117. Hand Shoveling. — Where possible the coal is dumped direct 
from the cars or wagons into bins located in front of the boilers. In 
such instances one man may handle the coal and ashes and attend to 
the water level of 200 horse power of boilers equipped with common 
hand-fired furnaces. With stoking and dumping grates 300 horse power 
may be controlled by one man and from 800 to 1000 horse power with 
chain-grate stokers. This refers of course to average good coal not 
too high in ash nor productive of much clinker. Sometimes the coal 
cannot be stored in front of the boilers but must be hauled by wheel- 
barrow, cart, or rail car. For distances over 100 feet and quantities 
over 20 tons per day the cost of handling the coal in this way may 
justify the installation of an automatic conveyor system. Hand-fired 
furnaces and manual handling of coal and ashes are usually associated 
with small plants of 500 horse power and under, but a number of large 
stations are operated in this way with apparent economy. A notable 
example is the new (1907) steam-power plant of the Wood Worsted 
Mill, Lawrence, Mass., in which 40 return tubular boilers are fired by 
hand. A tipcart with a capacity of one ton brings the coal a distance 
of 100 to 200 feet to the firing floor, and firemen shovel it on to the grate. 
Four men are stationed at the coal pile. One man drives two carts (one 
of which is being filled while the other is gone with its load), sixteen 
firemen attend to the furnaces, and two men dispose of the ashes. 



184 STEAM POWER PLANT ENGINEERING 

Most large plants, however, are equipped with conveying machinery, 
not so much because of the possible reduction in cost of operation, 
taking into consideration all charges fixed and operating, as because 
of the large and often unreliable labor staff which it dispenses with. 
Hand shoveling is sometimes necessary even with modern unloading 
devices on account of the freezing of coal in the cars. This is par- 
ticularly true of washed coals, and it is not unusual to have an entire 
car load solidly frozen. In this case it has to be picked and shoveled 
by hand, or the unloading tracks must be equipped with steam pipes 
and outfits for thawing purposes. A good man is capable of shoveling 
40 to 50 tons of coal in eight hours when unloading a car, provided 
it is only necessary to shovel the coal overboard. 

118. Bucket Conveyors. — One of the most common methods of 
automatically handling the coal from car to bunker is by means of an 
endless chain of traveling buckets. Many of the largest central stations 
in this country are equipped with such systems. The details of opera- 
tion are best illustrated by a few examples. 

Fig. 96 gives a diagrammatic arrangement of the link-belt over- 
lapping pivoted bucket carrier, and Fig. 97 illustrates its application 
to a typical boiler plant. Coal is discharged from the railway cars 
into a track hopper and from there delivered by a " feeding apron " 
into a crusher which reduces it to such a size as can be conveniently 
handled by the stokers. It is then discharged into a short bucket 
conveyor, which carries it to the main system of buckets, and it is 
elevated to the proper level and discharged into the overhead bunkers. 
The discharge is effected by special tripping devices which engage the 
buckets and turn them over. The ashes are dumped from the ash pit 
through a series of chutes into the lower run of buckets, by which they 
are elevated and discharged into the ash hopper alongside the coal 
bunkers. From the ash hopper the ashes discharge by gravity directly 
into the railway cars below. The system is operated by means of two 
motors, one driving the crusher and the other the main bucket system. 
The buckets are made of either sheet steel or malleable iron. 

In Fig. 96 the coal is fed to the crusher by the " reciprocating feeder," 
which is usually placed directly under the track hopper. The feeder 
consists of a heavy steel plate mounted on rollers and having a recip- 
rocating movement effected by a crank mechanism from the carrier. 
The amount of coal delivered depends upon the distance the plate moves, 
and this can be varied by changing the throw of the eccentric. The 
number of strokes corresponds to the number of buckets. Any size 
coal can be readily handled. When the distance from track hopper 
to carrier is so great that the reciprocating feeder is not practicable a 



COAL AND ASH-HANDLING APPARATUS 



185 




186 



STEAM POWER PLANT ENGINEERING 




I 

'3 

cr 



COAL AND ASH-HANDLING APPARATUS 



187 




188 



STEAM POWER PLANT ENGINEERING 




Fig. 99. Coal and Ash-Handling System in the Power House of the South Side 
Elevated Railway Company, Chicago. 



COAL AND ASH-HANDLING APPARATUS 189 

continuous or " belt " feeder is used to supply the crusher with fuel. 
The " equalizing gear " is designed to impart a pulsating motion to the 
driving sprocket wheel which will counteract the natural pulsation to 
which long pitch chains are subject, producing violent increase of the 
normal strain at frequent intervals. This is accomplished by driving 
the spur wheel with an eccentric pinion, causing the pitch line to describe 
a series of undulations corresponding to the number of sprockets on 
the chain wheel. Figs. 99 and 100 show the general arrangement of 
crusher and " cross conveyor " in the old portion of the South Side 
Elevated Power House, Chicago. 

A coal and ash system similar to the one illustrated in Fig. 97 for a 
plant consisting of eight 350-horse-power boilers will cost in the neigh- 
borhood of $8,000, completely installed. This does not include the 
cost of coal and ash bunkers. 

The Hunt conveyor, Fig. 101, while usually called a " bucket " 
conveyor, is in fact a series of cars connected by a chain, each having 
a body hung on pivots and kept in an upright position by gravity. The 
chain is driven by pawls instead of by sprocket wheels. The " buckets " 
are upright in all positions of the chain, consequently the chain can be 
driven in any direction. The change of direction of the chain is accom- 
plished by guiding the carriers over curved tracks. The chain moves 
slowly, and the capacity is governed by the size of the buckets. The 
ordinary size buckets carry two cubic feet of coal and move at a rate 
of fifteen buckets a minute, carrying about 40 tons per hour. Two 
methods of filling the buckets are employed, the " measuring " and the 
" spout filler." In the former each bucket is separately filled with a 
predetermined amount by a suitable " measuring feeder." In the 
latter the material is spouted in a continuous stream, necessitating 
the use of overlapping buckets to prevent spilling of the material. 
Fig. 102 shows an application of the Hunt system to the old plant 
of the Baltimore United Railways and Electric Company. 

Fig. 103 gives a sectional elevation of the coal and ash-handling 
machinery at the power plant of the Commercial National Bank Build- 
ing, Chicago. Underneath the sidewalk on the Clark Street side of the 
building is a coal-storage bin of 600 tons capacity, served with a bucket 
conveyor. One leg of the conveyor reaches down to a level below the 
track of the Illinois Tunnel Company. By this arrangement coal 
can be delivered either by cars in the tunnel or by wagons from the 
street. In taking coal from storage a gate at the lower extremity of 
the hopper is opened and the coal filling the buckets is elevated and 
tripped into any one of the screw conveyors leading from bucket con- 
veyor to boiler hopper. The ashes are shoveled from the ash pits into 



190 



STEAM POWER PLANT ENGINEERING 







i 



COAL AND ASH-HANDLING APPARATUS 



191 



j^^TT^ 




Fig. 101. Driving Mechanism of Hunt Conveyor. 




Fig. 102. Coal and Ash-Handling System at the Old Power House of the Baltimore 
United Railways and Electric Company. 



192 



STEAM POWER PLANT ENGINEERING 



cars running in a cross tunnel under the boiler floor, and by these cars 
are transferred to a dump at one side of the boiler room and discharged 
into Illinois Tunnel Company's cars for removal. 




Fig. 103. Bucket and Screw Conveyor at Commercial National Bank Building, 

Chicago, Illinois. 

119. Belt Conveyors. — The Robins belt conveyor, Fig. 104, consists 
essentially of a thick belt of the required width driven by suitable 
pulleys and carried upon idlers so arranged that the belt becomes 

trough-shaped in cross section. 
The belt is constructed of woven 
cotton duck covered with a 
special compound on the carry- 
ing side. The belt is thicker at 
the middle than at the edges, 
since the wear is greatest in a 
line along the center. The idlers 
are carried by iron or wooden 
framework, and are spaced from 
3 feet to 6 feet between centers 
on the troughing side according to the width of belt and the weight 
of the load. On the return side these distances range from 8 to. 12 feet. 




Ow 






Fig. 104. 



Guide Pulleys, Robins Belt 
Conveyor. 



COAL AND ASH-HANDLING APPARATUS 193 

High-speed rotary brushes with interchangeable steel bristles prevent 
wet, sticky material from clinging to the belt. Automatic tripping 
devices placed at the proper points cause the material to be discharged 
where it is needed. The trippers consist essentially of two pulleys, 
one above and slightly in advance of the other, the belt running over 
the upper and under the lower one, the course of the belt resembling 
the letter S. The material is discharged into chutes on the first down- 
ward turn of the belt. The trippers may be movable or fixed, single or 
in series. Movable trippers are used when it is desired to discharge 
the load evenly along the entire length, as, for instance, in a continuous 
row of bins, while fixed trippers are employed where the load is to be 
discharged at certain and somewhat separated points. The movable^ 
trippers are made in two forms, " hand-driven " and "automatic." In 
the former they are moved from point to point by means of a hand 
crank. The " automatic " tripper is propelled by the conveying belt 
through the medium of gearing. • It reverses its direction automatically 
at either end of the run, and travels back and forth continuously dis- 
tributing its load. It can be stopped, reversed, or made stationary at 
will. The most notable installations of this system are at the 96th Street 
station and the Kingsbridge station of the Metropolitan Street Railway 
Company, New York City. 

120. Elevating Tower, Hand-Car Distribution. — Fig. 105 illus- 
trates the coal and ash-handling installation at the Aurora and Elgin 
Interurban Railroad power house, Batavia, 111. Coal is delivered to 
the plant by railroad cars which dump directly into coal hoppers 
located inside a steel structure running the entire length of the building 
and spanned by two railroad tracks. There are 18 hoppers constructed 
of 17-inch brick walls fitted with steel -plate bottoms. Subdividing 
the storage space in this manner makes it possible to carry different 
grades of coal, prevents the spreading of fire, and affords a simple con- 
struction for the support of the railroad tracks. The basement of the 
boiler room extends underneath the hoppers, and two lines of narrow- 
gauge tracks are imbedded in the concrete floor. Turntables at the 
center facilitate the switching of cars to the elevators which rise through 
the boiler room close to the chimney. The cars, of one ton capacity 
each, are of special construction, with roller-bearing axles and a com- 
bined ratchet lift and friction dump. The filled cars are pushed from 
underneath the hoppers to two elevators which lift them to the line of 
tracks supported overhead across the boiler fronts. They are then 
pushed to the hoppers suspended above the boiler setting and the coal 
is dumped. These hoppers have a capacity of six tons each. From 
the hoppers the coal is fed to the stoker by an ordinary down spout. 



194 



STEAM POWER PLANT ENGINEERING 



The ashes fall from the stokers into an ash pit, from which they may be 
discharged into ash cars. The ash cars are elevated to a set of tracks 
running at right angles to the main tracks, and are transferred to ash 
bins located directly over the coal bins. Coal and ashes are weighed 




GRADE 



TRACK TO ELEVATOR 



Fig. 105. Coal and Ash-Handling System at the Power House of the Aurora and Elgin 
Interurban Railway, Batavia, 111. 



in the small cars. There are ten boilers in this plant and four men are 
required to handle the coal and ashes. The entire coal and ash-handling 
system cost about $10,000, and. the cost of handling the coal and 
ashes is approximately 4 cents per ton. This does not include wages 
of firemen or water tenders. 



COAL AND ASH-HANDLING APPARATUS 



195 



121. Overhead Storage, Bucket Hoist. — Fig. 106 gives a general 
view of the coal-handling plant of the Depot Street power house of the 
Cincinnati Traction Company. This installation is a good example of 
an application of the " overhead storage gravity feed " system to an 
existing plant without interfering in any way with its operation. The 
system consists essentially of a receiving pit below the car tracks from 
which the coal is hoisted to a series of overhead bins. The coal storage 
is outside the boiler house in an independent structure. The bins 
are of steel framework with concrete floors, and are sufficiently elevated 



Fig. 




PIT CAPACITY 
SO TONS 



Coal and Ash-Handling System at the Depot Street Power House of the 
Cincinnati Traction Company. 



to spout coal easily to the stoker magazine. The total capacity of the 
overhead bins is about 1,600 tons. The four bins or receiving pits have 
a capacity of 50 tons each, or approximately one car load, and are so 
situated that all four may be filled simultaneously without shifting the 
train. The coal-handling apparatus consists of a one-ton self-filling 
bucket operated on a three-motor electric crane running on rails at 
the top of the storage bins. The coal is hoisted from the receiving pit 
through suitable shafts in the bin structure and dumped into the over- 
head hoppers. The maximum capacity of the hoist is 50 tons per hour. 
The labor required to handle the coal from car to bins is performed by 
one man working five hours per day and an assistant engaged a small 
part of the time to dump cars, clean hoppers, etc. The average daily 
coal consumption is approximately 200 tons. The total cost of the 



196 STEAM POWER PLANT ENGINEERING 

equipment was about $18,000 for the bins complete and $4,500 for the 
coal-handling crane. The cost of handling the coal and ashes is approxi- 
mately 1.5 cents per ton of coal. Including all charges fixed and operat- 
ing the total cost of handling the coal is about 3.5 cents per ton. This 
does not include wages of firemen or water tenders. 

122. Elevating Tower, Cable-Car Distribution. — The coal and ash- 
handling system of the new turbine power plant of the Detroit Edison 
Company, Fig. 107, is a typical example of a large station equipped 
with elevating tower and cable-car distributers instead of the usual 
bucket conveyor. The system consists essentially of a lofty steel 
tower in which are housed at various levels a track receiving hopper, 
crushing rolls and screens, weighing hopper, hoisting apparatus, etc., 
and a small cable railway for delivery to the bunkers. The railroad 
coal cars enter the tower on an elevated trestle 18 feet above grade, 
below which is a track receiving hopper. A two-ton " tub hoist " is 
filled with coal from the bottom of the receiving hopper and elevated 
to a 20-ton bin at the top, 120 feet above ground level. This bin has 
a grille bottom at one side and under the outlet a heavy duty coal crusher, 
thus allowing the fine coal to screen through directly while all the 
larger lumps are automatically -delivered to the crusher. The hopper 
beneath this delivers to the revolving screen, which sorts the slack into 
one bin below and the nut coal into the other. From the two bins the 
small cable cars are filled for dumping into the desired bunkers over 
the boiler rooms. The cars are arranged for automatic dumping by 
means of adjustable trips which may be located at any point. The 
object of separating the nut coal and slack is to burn the latter during 
light or medium loads, keeping the former for heavy loads and " peak " 
overloads. The down spouts are double, with a valve in each branch 
operated from the floor, so that either grade of fuel may be drawn out 
at any time and in any proportion desired. The entire system has a 
capacity of from 50 to 75 tons of coal per hour and is driven by steam 
engines, with the exception of the revolving screen which is motor 
driven. The ash-handling system consists of brick-lined concrete hop- 
pers underneath each pair of stokers which discharge their contents 
by gravity into the small cars operated on the track system in the 
boiler-house basement. 

When handling 275 tons per day of 24 hours the cost of operation 
is approximately 12.5 cents per ton from coal car to ash car, including 
wages of firemen and water tenders. 

123. "Vacuum" Ash Conveyor. — Fig. 108 gives a diagrammatic 
arrangement of a recently patented ash-conveying system depending 
upon the velocity of a column of air for moving the ashes. The system 



COAL AND ASH-HANDLING APPARATUS 



197 



^Os. 



RECEIVING 
BIN 




Fig. 107. Coal and Ash-Handling System at the Power House of the 
Detroit Edison Company, 



198 



STEAM POWER PLANT ENGINEERING 



is simple in operation and low in first cost. One end of special cast- 
iron header F leads to the ash pits of the various boilers by means of 
branch tubes, and the other end is connected with a sealed separating 
chamber A. Each branch pipe is fitted with simple circular openings 
directly underneath each ash-pit door for admitting ashes and which 
are kept covered except when in operation. Exhauster E creates a 
partial vacuum in chamber A and draws in air at a high velocity from 
the opening in the ends of the branch pipes. Ashes raked into the 




Fig. 108. Diagrammatic Arrangement of the "Vacuum" Ash-Handling System. 



pipes through the openings are caught by the rapidly moving column 
of air and forced into chamber A. The ashes fall to the bottom and are 
fed into the main ash pit by a slowly revolving ash valve B. Air and 
dust are withdrawn from the top of the separator chamber through 
pipe G and discharged to the stack or to waste. A spray is introduced 
into pipe F to reduce dust. The process is a continuous one, and the 
ashes may be completely removed from the ash bin without interfering 
with the operation of the exhauster. In a later construction the ash 



COAL AND ASH-HANDLING APPARATUS 



199 




200 



STEAM POWER PLANT ENGINEERING 




COAL AND ASH-HANDLING APPARATUS 201 

bin and separating chamber are included in one chamber, thus doing 
away with the revolving ash valve and the small motor operating it. 
In this latter design the bin is never completely empty, a certain depth 
of ashes being maintained to seal the bottom at all times. 

At the Armour Glue Works, Chicago, 111., this system is applied to a 
boiler plant of thirteen boilers, aggregating 4,800 horse power, and cost, 
completely installed, $5,600. As originally installed the separating 
chamber had a volume of about 35 cubic feet and the suction intake was 
placed 58 feet above the ash-pit level. The revolving ash valve made 
about 13 r.p.m., and was driven by a one-horse-power motor. In the 
present installation the separating chamber and motor-operated ash 
valve are dispensed with and the discharge pipes lead directly into the 
main ash bin, which has a capacity of 60,000 pounds of wet ashes and is 
constructed of five-sixteenths -inch sheet iron. The exhauster (a 30-foot 
Root blower) has a capacity of about 8,000 cubic feet per minute at 
265 r.p.m., and is driven by a 75-horse-power motor. Under normal 
conditions of operation the motor requires 50 horse power when deliver- 
ing 250 pounds of ash per minute, and the vacuum on the suction side 
of the exhauster is 3.3 inches of mercury. The pipe from the ash bins 
to the separating chamber is 10 inches in diameter and is constructed 
of extra heavy chilled cast-iron pipe. The piping from the separating 
chamber to exhauster and to stack is 22 inches in diameter and is con- 
structed of number 16 and number 20 galvanized iron. The ashes are 
raked by hand from the ash pits to the suction openings of the branch 
pipes, and are handled dry, the dust being taken along with the ashes. 
Elbows are soon worn out by the abrasive action of the ashes, and tees 
are used instead, since the accumulation in the " dead " end receives the 
impact and takes up the' wear. The cost of handling the ashes in this 
installation is approximately 7 cents per ton. 

124. Cost of Handling Coal and Ashes. — In large stations where a 
number of men are employed to handle coal and ashes only it is a simple 
matter to divide the cost of handling into the various stages, thus : 

1. Cost of unloading cars or barges. 

2. Cost of conveying coal to bunkers. 

3. Cost of feeding coal to furnace. 

4. Cost of removing ashes. 

These costs are usually expressed in cents or dollars per ton of coal 
burned, or in terms of cents or dollars per horse power hour or kilo- 
watt hour of main prime mover output. Item number 3 is oftentimes 
included under " boiler-room attendance " and items 1, 3, and 4 under 
" coal and ash handling." Not infrequently all four items are included 
under " attendance." So much depends upon the character of stokers 



202 STEAM POWER PLANT ENGINEERING 

and furnace, size of boilers, and the like, that general figures on the cost 
of handling the coal and ashes are of little value unless accompanied by 
a description of the equipment. For the sake of general comparison 
the most satisfactory method of expressing the cost is in dollars per ton 
of coal from coal car to ash car. This includes wages of coal and ash 
passers, repair men, and boiler tenders. In small stations the coal 
and ash handling is done by the boiler tenders, in which case it is 
impracticable to separate the items mentioned above, and the cost is 
ordinarily included under attendance. An average figure for handling 
coal by barrow and shovel is not far from 1.6 cents per ton per yard 
up to the distance of five yards, then about 0.1 cent per ton per yard 
for each additional yard. With automatic conveyors the operating 
cost, not including wages of firemen and water tenders, varies with the 
size of plant and the type of conveyor, and ranges anywhere from a 
fraction of a cent per ton to four or five cents per ton. The larger the 
plant and the greater the amount of coal handled the lower will be the 
cost per ton. In comparing the relative costs of manual and automatic 
handling, fixed charges of at least 15 per cent of the first cost of the 
mechanical equipment should be charged against the latter in addition 
to the cost of operation. In large central stations equipped with stokers 
and conveyors and consuming 200 tons or more of coal in twenty-four 
hours, the cost of handling the coal from coal car to ash car, including 
wages of firemen and water tenders, will range between 10 cents and 18 
cents a ton. 

125. Coal Hoppers. — Fig. 109 shows a front and side elevation of 
a typical set of stationary weighing hoppers as applied to the boilers 
of the Quincy Point power plant of the Old Colony Street Railway 
Company, Quincy Point, Mass. Each battery of boilers is provided 
with an independent set of hoppers. The bottoms of the overhead 
coal bunkers lead into the small hoppers A, A. The operation of any 
single weighing hopper is as follows: Coal is fed from the overhead 
bunkers to weighing hopper H by means of valve V. The weight of 
coal in the weighing hopper is transmitted by a system of levers and 
knife edges to the inclosed scale beam / and noted in the usual way. 
The weighed charge of coal is then admitted to the down spout S by 
means of valves similar to those at V. 

Although separate weighing hoppers for each battery, as illustrated 
in Fig. 109, offer many advantages, they are quite costly and it is not 
unusual to install one or more large weighing hoppers mounted on 
overhead traveling carriages so that one may supply a number of 
boilers (Fig. 110). At the Armour Glue Works, Chicago, the coal supply 
is stored in one large overhead bunker of 1000 tons capacity. A five- 



COAL AND ASH-HANDLING APPARATUS 



203 




Fig. 109. Stationary Coal Weighing Hoppers. 




Fig. 110. Traveling Coal Hoppers. 



204 



STEAM POWER PLANT ENGINEERING 



ton motor-driven traveling hopper receives its supply from this central 
bunker and delivers it to the various boilers. One man operates the 




Fig. 111. Common Slide Coal Valve. 



Fig. 112. Simplex Coal Valve. 



traveling hopper, tends to the coal valves, and supplies all boilers with 
coal. 

Weighing hoppers are sometimes made automatic; that is, the opening 
and closing of valves, feeding of coal, and recording of weight are auto- 
matically performed by the 
weight of the coal itself. The 
scale is set for discharges of a 
certain weight and continues 
to discharge this amount auto- 
matically. In the few plants 
which are equipped with auto- 
matic weighing hoppers the 
capacity of the hopper is 
approximately 100 pounds per 
discharge. These hoppers are 
necessarily more complicated 
and more costly than the 
ordinary weighing hoppers, 
and it is a question whether 
the advantages offset the 
extra first cost and main- 
tenance charges. A small automatic hopper of 100 pounds discharge 
capacity costs approximately $400 as against $250 for the ordinary 
weighing device. 




Fig. 113. Duplex Coal Valve. 



COAL AND ASH-HANDLING APPARATUS 



205 



126. Coal Valves. — Figs. Ill to 115 illustrate the principles of a few 
well-known coal valves. They may be conveniently grouped into two 
classes according to the location of the coal pocket: (1) those drawing 
the coal from overhead bunkers and (2) 
those drawing from the side of a bin. In 
the first class come the simple slide valve, 
the simplex and duplex rotating valve. In 
the latter are the flap valve and the 
rotating valve. They are made in various 
sizes and designs, but those illustrated are 
examples of the most common types. The 
simple slide valve, Fig. Ill, is applicable 
only to small size coal and to small spouts, 
since coarse or lump coal may get in the way 
and prevent proper closing. The simplex 
valve, Fig. 112, consists of a rotating jaw 
actuated by a lever. There are no rubbing 
surfaces, and the jaws cut through the 
material without jamming. The duplex 





Fig. 



114. Common 
Coal Valve. 



Flap 



Fig. 115. " Seaton " Coal Valve. 



valve, Fig. 113, consists 
of two rotating jaws con- 
nected to a common 
actuating lever. The 
jaws move simultane- 
ously, so that even a par- 
tially open valve delivers 
the coal centrally. When 
closing the valve the flow 
is gradually stopped by 
the decreasing width of 
the opening and there is 
but little resistance to 
The largest valve can easily be operated 



the movement of the jaws, 
by hand. 

The flap valve, Fig. 114, is the simplest form for drawing coal from a 
side bin. It consists merely of an iron flap hinged to the bottom of 



206 STEAM POWER PLANT ENGINEERING 

the chute. The valve is lowered to let the coal run over its top and is 
raised to stop the flow. It cannot be clogged or get jammed in closing. 
The flap is raised and lowered by a simple lever. For very large bins, 
where the valves are to be opened and closed frequently, the " Seaton " 
valve, Fig. 115, is usually preferred. This valve consists of two jaws 
EE', and TT' pivoted to suitable framework at and actuated by 
lever A. The valve is shown fully closed. Raising lever A causes the 
cut-off blade EE / to rotate about and permits the coal to flow 
through the space between the edge of the jaw E and the end of the 
chute. The rate of flow is regulated by the width of this opening. The 
cut-off blade does not reach a stop, hence there is no possibility of a 
lump of coal getting in the way and preventing the prompt closing of 
the valve. 

Coal and Ash-Handling Installations : Commonwealth Edison, Chicago, Power, 
Dec, 1906, p. 718. Boston Elevated, Elec. World, Sept. 7, 1901, p. 396. Inter- 
borough Rapid Transit Co., New York, Elec. World, Feb. 4, 1905, p. 264; Engr., 
U.S., May 15, 1904, p. 337; Eng. News, Jan. 14, 1904, p. 41. Waterside Station, 
New York, Edison Co., Eng. Rec, Sept. 9, 1905, p. 287. Detroit Edison Co., Eng. 
Rec, Oct., 1905, p. 396. Brooklyn Rapid Transit Co., St. Ry. Jour., Sept. 23, 
1905, p. 435. N. Y. C. and H. R. R., St. Ry. Jour., Nov. 11, 1905, p. 876. Aurora, 
Elgin and Chicago Ry., Eng. Rec, Feb. 7, 1903, p. 153. Missouri River Power 
Co., Eng. News, Oct. 19, 1905, p. 403. Brooklyn Edison Co., Gold St. Station, 
Elec World, June 15, 1907. 

Hoisting and Conveying Machinery: Pro. A.S.M.E., June, 1908. 



CHAPTER VII. 

CHIMNEYS. 

127. Chimney Draft. — Draft produced by a chimney depends upon 
so many conditions and involves such a large number of variables that 
empirical methods of proportioning, based upon actual performances, 
are more to be relied upon than theoretical calculations. Draft is 
due to the difference in the weight of the column of hot light gases 
in the stack and that of the cooler and heavier surrounding atmos- 
phere, the latter tending to flow into the base and thereby force the 
lighter gases out the top of the stack. The commonly accepted theory 
of chimney draft is based upon Peclet's hypothesis that the flow through 
the furnace flues and chimney may be represented by the equation 

h== £Tl( l+ G + — )> < 27 > 

64.4 \ m ) 

in which 

h = the head of fluid producing the flow, feet. 

u = velocity of the gases in the chimney, feet per second. 

G = a coefficient to represent the resistance to the passage of air 

through the coal. 
I = total length of the path of the gases, feet. 
m — area of cross section divided by the perimeter. 
/ = a coefficient depending upon the nature of the surfaces over 
which the gases pass. 

From experiments on chimneys and boilers Peclet gives in connection 
with this theory the following values of coefficients G and / : 

G - 12, / = 0.012, 

on the basis of 20 to 24 pounds of coal burned per square foot of grate 
surface per hour. On account of the variation in practice of the factors 
u, f, and G and the difficulty of determining them engineers prefer to 
use the modified formulas given further on. 

The difference of pressure, or intensity of draft may be expressed 
theoretically, ignoring friction, as follows : 

207 



208 STEAM POWER PLANT ENGINEERING 

Let H — height of chimney in feet. 

T = absolute temperature of the freezing point, degrees F. 
T x — absolute temperature of the gases in the chimney. 
T 2 = absolute temperature of the outside air. 
P = average atmospheric pressure. 
P 2 = observed atmospheric pressure. 

W = weight of a cubic foot of air at 32 degrees F. and pressure P. 
W x = weight of a cubic foot of chimney gas at 32 degrees F. 
and pressure P. 

Then the weight of a cubic foot of hot gas in the chimney will be 

W * %■'■£[ (28) 

and the weight of a cubic foot of cold air outside will be 

w 7? • # ■ (29) 

The weight of a column of hot gas H feet high and one foot square 
will be 

W X H^. |L. (30) 

Similarly the weight of the cold-air column will be 

WH ^-Y 2 (31) 

and the difference in pressure or the intensity of draft will be 

where D is in pounds per square foot. 

By making P = P 2 = 14.7, T = 493, W = 0.0807, W. x = 0.084, 
and D x = pressure in inches of water (D t = 0.192 D), equation (32) 
assumes the familiar form 

JW( 7 -f-?f). 03) 

By assuming W = W x = 0.081 and P = 14.7, equation (32) may 
be written 



A=0.52tfP 2 (i-I-) 



(34) 



This latter form is ordinarily used where the atmospheric pressure 
differs considerably from that at sea level, as at high altitudes. Table 23 
gives the density of air and chimney gases at various temperatures. 



CHIMNEYS 



209 



Example : Required the maximum theoretical draft obtainable from 
a chimney 150 feet high, atmospheric pressure 14.7 pounds per square 
inch, temperature outside air 60 degrees F., temperature chimney 
gases 550 degrees F. 

Here H = 150, T 2 = 461 + 60 = 521, T x = 461 + 550 =1011. 

Substituting these values in equation (34), 

D x = 150 (— - — ) = 1.02 inches of water, 

which is about 25 per cent greater than the draft actually obtained, 
and represents the maximum possible under the given conditions, 
neglecting the resistance offered by the chimney and the pressure 

TABLE 23. 

DENSITY AND SPECIFIC VOLUME OF AIR AND CHIMNEY GASES AT 
VARIOUS TEMPERATURES. 



Air. 


Chimney Gases. 


t 


5 


V 


d 


t 


d 


t 


d 


t 


d 





11.581 


.935 


.086353 


200 


.06334 


430 


.04695 


660 


.03730 


5 


11.706 


.945 


.085424 


210 


.06239 


440 


.04643 


670 


.03697 


10 


11.832 


.955 


.084513 


220 


.06147 


450 


.04592 


680 


.03665 


15 


11.931 


.965 


.083623 


230 


.06058 


460 


.04542 


690 


.03633 


20 


12.085 


.976 


.082750 


240 


.05971 


470 


.04493 


700 


.03602 


25 


12.211 


.986 


.081895 


250 


.05887 


480 


.04445 


710 


.03571 


30 


12.337 


.996 


.081058 


260 


.05805 


490 


.04398 


720 


.03540 


32 


12.387 


1.000 


.080728 


270 


.05726 


500 


.04353 


730 


.03511 


35 


12.463 


1.006 


.080238 


280 


.05648 


510 


.04308 


740 


.03481 


40 


12.589 


1.016 


.079434 


290 


.05573 


520 


.04264 


750 


.03453 


45 


12.715 


1.026 


.078646 


300 


.05499 


530 


.04221 


760 


.03424 


50 


12.841 


1.037 


.077874 


310 


.05428 


540 


.04178 


770 


.03396 


55 


12.967 


1.047 


.077117 


320 


.05358 


550 


.04137 


780 


.03369 


60 


13.093 


1.057 


.076374 


330 


.05290 


560 


.04096 


790 


.03342 


62 


13.144 


1.061 


.076081 


340 


.05224 


570 


.04056 


800 


.03316 


65 


13.220 


1.067 


.075645 


350 


.05159 


580 


.04017 


900 


.03072 


70 


13.346 


1.077 


.074930 


360 


.05096 


590 


.03979 


1000 


.02861 


75 


13.472 


1.087 


.074229 


370 


.05035 


600 


.03942 


1100 


.02678 


80 


13.598 


1.098 


.073541 


380 


.04975 


610 


.03905 


1200 


.02516 


85 


13.724 


1.108 


.072865 


390 


.04916 


620 


.03869 


1300 


.02373 


90 


13.851 


1.118 


.072201 


400 


.04859 


630 


.03833 


1400 


.02245 


95 


13.976 


1.128 


.071550 


410 


.04803 


640 


.03798 


1500 


.02131 


100 


14.102 


1.138 


.070910 


420 


.04749 


650 


.03764 


1800 


.01848 


110 


14.354 


1.159 


.069665 










2000 


.01698 















d = density, pounds per cubic foot. 

t = temperature, degrees F. 

s = specific volume, cubic feet per pound. 

v = comparative volume, volume at 32° = 1. 

Density of chimney gas taken 0.085 pound per cubic foot at 32° F. and 29.92 
inches of mercury. 

(Rankine, " Steam Engine," gives the density at 32° F. as varying from 0.084 to 
0.087.) 



210 



STEAM POWER PLANT ENGINEERING 



TABLE 24. 

THEORETICAL DRAFT PRESSURE IN INCHES OF WATER. CHIMNEY 
100 FEET HIGH. 1 



Temp. 


Temperature of the External Air - 


- Barometer, 14.7 Pounds per Square Inch. 2 


in the 
























Chim- 
























ney. 


0° 


10° 


20° 


30° 


40° 


50° 


60° 


70° 


80° 


90° 


100° 


200 


.453 


.419 


.384 


.353 


.321 


.292 


.263 


.234 


.209 


.182 


.157 


220 


.488 


.453 


.419 


.388 


.355 


.326 


.298 


.269 


.244 


.217 


.192 


240 


.520 


.488 


.451 


.421 


.388 


.359 


.330 


.301 


.276 


.250 


.225 


260 


.555 


.528 


.484 


.453 


.420 


.392 


.363 


.334 


.309 


.282 


.257 


280 


.584 


.549 


.515 


.482 


.451 


.422 


.394 


.365 


.340 


.313 


.288 


300 


.611 


.576 


.541 


.511 


.478 


.449 


.420 


.392 


.367 


.340 


.315 


320 


.637 


.603 


.568 


.538 


.505 


.476 


.447 


.419 


.394 


.367 


.342 


340 


.662 


.638 


.593 


.563 


.530 


.501 


.472 


.443 


.419 


.392 


.367 


360 


.687 


.653 


.618 


.588 


.555 


.526 


.497 


.468 


.444 


.417 


.392 


380 


.710 


.676 


.641 


.611 


.578 


.549 


.520 


.492 


.467 


.440 


.415 


400 


.732 


.697 


.662 


.632 


.598 


.570 


.541 


.513 


.488 


.461 


.436 


420 


.753 


.718 


.684 


.653 


..620 


.591 


.563 


.534 


.509 


.482 


.457 


440 


.774 


.739 


.705 


.674 


.641 


.612 


.584 


.555 


.530 


.503 


.478 


460 


.793 


.758 


.724 


.694 


.660 


.632 


.603 


.574 


.549 


.522 


.497 


480 


.810 


.776 


.741 


.710 


.678 


.649 


.620 


.591 


.566 


.540 


.515 


500 


.829 


.791 


.760 


.730 


.697 


.669 


.639 


.610 


.586 


.559- 


.534 


550 


.863 


.828 


.795 


.762 


.731 


.700 


.671 


.644 


.618 


.593 


.585 


600 


.908 


.873 


.839 


.807 


.776 


.746 


.717 


.690 


.663 


.638 


.613 



1. For any other height multiply the tabular figure by ^-r t where H is the height in feet. 

p. 

2. For any other pressure multiply the tabular figure by , where P is the barometric pres- 
sure in pounds per square inch. 



required to impart velocity to the gases. Table 24 has been computed 
from formula (34), and gives the maximum theoretical draft in a chim- 
ney 100 feet high for different flue-gas temperatures. 

The intensity of draft required to produce best results depends upon 
the kind and condition of fuel, the thickness of fire, character of grate, 
and resistance of the breeching, tubes, baffles, dampers, etc. As stated 
above, the loss of draft in the chimney proper approximates 20 per cent of 
the total, that in the breeching is taken as 0.1 inch per 100 feet of flue, 
and 0.05 inch for each right-angle bend; the loss in the boiler varies from 
0.3 to 0.6 inch, depending upon the type;* the loss in the furnace varies 
between wide limits, and depends upon the kind of fuel and the rate of 
combustion. The curves in Fig. 116 compiled by the Stirling Company 
and published in their book " Stirling'' give the furnace drafts necessary to 
burn various kinds of fuels at different combustion rates, and give an idea 
of the influence of the character of the fuel and the rate of combustion. 
* Specific figures may be obtained from the manufacturers. 



CHIMNEYS 



211 







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212 STEAM POWER PLANT ENGINEERING 

Example : Determine the probable draft necessary to burn 30 pounds 
bituminous run of mine per hour per square foot of grate when the out- 
side air is 60 degrees F., the temperature of the chimney gases 550 
degrees, and the flue is 100 feet long, with two right-angle bends. 

The losses will be divided approximately as follows : 

Inch. 

Loss in furnace (from curves in Fig. 116) 0. 17 

Loss in boiler (average) 0. 40 

Loss in flue, 100 feet at 0.10 per 100 . 10 

Loss in turns, 2 X 0.05 .10 

6777 

Since the loss in the chimney alone approximates 20 per cent of the 
total, 0.77 -T- .80 = 0.96 will be the theoretical draft necessary. From 
equation (33), 

Substituting for the given values of D lf T lf and T 2 in above equation, 

°- 96 - H fer ~ ion] ' 

From which, 

H = 142, height of stack necessary to produce a draft of 0.17 inch 
in the furnace. 

Table 25 gives the results of a test of a 100-foot unlined steel chimney, 
showing the variation in draft at different points in the stack. 

Theory of Chimney Draft : Power, Oct., 1896, p. 18, Dec, 1898, p. 20, March, 
1906, Feb., 1900, p. 12; Engr. U.S., Jan. 15, 1903, May 15, 1902, p. 313; Trans. 
A.S.M.E., 11-451, 762, 772, 974, 984; Elec. Rev., Lond., Oct. 14, 1904. 

128. Chimney Formulas. — Rational methods of determining the 
height and area of chimneys being cumbersome and unwieldy and of 
doubtful value for practical use, the various empirical formulas outlined 
in Table 26 are quite commonly used. They give good results within 
the limits of the assumptions upon which they are based, but otherwise 
may lead to absurd results, their applicability depending largely upon 
the available data covering the various losses with the particular kind, 
quality, and condition of coal, and conditions of operation. Occasionally 
practical and local considerations fix the height of the stack irrespective 
of theoretical deductions. The logical procedure is to determine first 
the height of chimney necessary to produce the draft at the desired maximum 
rate of combustion, and then to proportion the area by such formulas as 
(2), (4), or (5), to suit the quantity of fuel to be burned. 



CHIMNEYS 



213 



The following heights have been found to give good results in plants 
of moderate size: 

Feet. 

With free-burning bituminous coal 80 

With anthracite, medium and large sizes 100 

With slow-burning bituminous 120 

With anthracite pea 130 

With anthracite buckwheat 150 

With anthracite slack 175 



TABLE 25. 

CHIMNEY DRAFT. 

Test of a 100-Foot Unlined Steel Chimney 3 Feet in Diameter at Massachusetts 
Institute of Technology. (Peabody & Miller, "Steam Boilers," p. 121.) 



Over the grate 

At the bridge wall 

Half-way between bridge and back end 

of boiler 

At the back end of boiler 

In uptake near boiler 

In stack 34 feet above grate 

In stack 51 feet above grate 

In stack 68 feet above grate 

In stack 85 feet above grate 



Draft, Inches 
of Water. 



Maximum. Minimum 



24 
382 



0.410 
0.354 
0.572 
0.440 
0.334 
0.216 
0.122 



0.218 
0.372 

0.374 
0.334 
0.543 
0.414 
0.312 
0.168 
0.086 



Temperature, Fah- 
renheit. 



Maximum. Minimum 



403 
396 
380 
370 
345 



389 
374 
368 
354 
314 



The chimney serves two 80-horse-power boilers. During test one was 
banked and the combustion at the grate of the working boiler was 19.8 
pounds per square foot of grate surface per hour. Coal burned per 
hour 590 pounds. 

For plants of 800 horse power or more the height of stack should 
never be less than 150 feet, regardless of the kind of coal used. 

Referring to Table 26, formulas (1), (2), (6), (7), and (9) are based upon 
a fuel consumption of 13 to 15 pounds of anthracite and 22 to 26 
pounds of bituminous coal per square foot of grate area per hour. In 
formulas (3), (4) and (9), the diameter is dependent solely upon the 
quantity of coal burned per hour and the height is determined mainly 
by the rate of combustion per square foot of grate. The results accord 
well with practice. With western coals formula (3) gives results rather 
too large and the constant should be 120 instead of 180. Formula (5) is 



214 STEAM POWER PLANT ENGINEERING 

perhaps the most used and has met with much approval. It is based on 
the assumptions that 

1. The draft of the chimney varies as the square root of the height. 

2. The retardation of the ascending gases by friction may be con- 
sidered due to a diminution of the area of the chimney or to a lining of 
the chimney by a layer of gas which has no velocity and the thickness 
of which is assumed to be 2 inches. Thus, for square chimneys, 

E=&-^- -A-\V1, (35) 

and for round chimneys, 



E =l( V -*-£) =^- 0-591 x/A 



(36) 



For simplifying calculations the coefficient of V A may be taken as 
0.6 for both square and round chimneys, and the formula becomes 

E = A- 0.6 VI (37) 

3. The horse-power capacity varies as the effective area E. 

4. A chimney should be proportioned so as to be capable of giving 
sufficient draft to permit the boiler to develop much more than its rated 
power in case of emergencies or to permit the combustion of 5 pounds 
of fuel per rated horse power per hour. 

5. Since the power of the chimney varies directly as the effective 
area E and as the square root of the height H, the formula for horse 
power for a given size of chimney will take the form 

H.P. = CE Vlf, (38) 

in which C is a constant, found by Mr. Kent to be 3.33, obtained by 
plotting the results from numerous examples in practice. 
The formula then assumes the form 

H.P. = 3.33 eVh ■ (39) 

H.P. = 3.33 (A - 0.6 VA) VW, (40) 



or 

from which 



H 



««£)■: m 



Table 27 has been computed from equation 5, Table 26. 
Many engineers simply adopt the following proportions : 
Internal area of chimney at top, one-seventh grate area for bitumi- 
nous coal. 





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STEAM POWER PLANT ENGINEERING 



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CHIMNEYS 21 7 

Internal area of chimney at top, one-ninth grate area for anthracite 
coal. 

Example : Determine the area and diameter of a stack for a 2000- 
horse-power plant to operate under the following conditions: Rated 
load 2000 horse power; maximum overload 40 per cent of rated; flue 
150 feet long, with one right- angle bend; average rate of combustion 20 
pounds of bituminous coal per square foot of grate surface per hour; 
atmospheric temperature 60 degrees F.; flue-gas temperature at over- 
load 600 degrees F.; coal burned per boiler horse power, 4 pounds. 

With modern types of steam engines or turbines an overload of 40 
per cent has little effect on the economy of the prime mover, and the 
boiler efficiency is but slightly reduced, but an additional allowance of 
25 per cent should be made in estimating the overload combustion rate. 

The maximum rate of combustion then will be 

pounds per square foot of grate surface per hour. 

The draft required at the point where the flue enters the chimney, 
considering the various losses, will be found as follows : 

Inch. 

Furnace (see curves. Fig. 116) 0.3 

Boiler ..0.4 

Flue, 150 feet at 0.1 inch per 100 feet 0. 15 

Turns, 1 at 0.05 0.05 

CUT 

From formula (33), 

•>-'<!%-'■$■ 

Substituting the following values : 

T = 60 + 461 = 521 ; 7\ = 600 + 461 = 1061 



0.9 



D 1 = theoretical draft = -^- = 1.12 inch. 

■ (7M 7.95\ 
1,12 - H {-521-T0Gl)' 



whence the necessary height of stack is 

H = 160 feet (approximately). 



218 STEAM POWER PLANT ENGINEERING 

Substituting the value of H in Kent's formula, the effective area is 
found to be 

E = ,— 1 = — — 7= = 47.5 square feet, 

VH Vl60 

corresponding to an actual diameter of 93 inches. 

Chimney Design: Engr. U.S., Jan. 1, 1907, p. 81, Feb. 1, 1907, p. 174; Am. 
Elecn., March, 1904, p. 123; Eng. Rec, April 18, 1900, July 27, 1901, pp. 52, 82, 
Sept. 21, 1901, p. 271, March 1, 1902, p. 195, July 19, 1902, p. 60, Nov. 22, 1902, 
p. 495, May 5, 1906, p. 549; Power, Jan., 1902, p. 12, Nov., 1902, p. 29, Dec. 1, 
1903, p. 66, Dec, 1901, p. 570, Dec, 1905, p. 745; Am. Elecn., Dec, 1901, p. 570; 
Cassier's Mag., July, 1904, p. 341, Feb., 1906, p. 267; Engr. U.S., Oct. 1, 1899, 
p. 222, Sept. 1, 1904, p. 591, June 15, 1905, p. 403; Eng. Mag., Oct., 1899, p. 124, 
Trans. A.S.M.E., 11-451; Eng. News, July 20, 1905, p. 64; Engr. U.S., Jan. 1, 
1907, p. 91. 

129. Height of Chimneys for Boilers using Oil Fuel. — Experi- 
mental data relative to chimneys for boilers using oil fuel are rather 
meager, and discordant, but a study of a number of recent installations 
seems to indicate that the area need not exceed 50 per cent of that 
required by the same boiler using bituminous coal. A height 80 to 
90 feet above the grate usually affords sufficient draft to force the 
boilers 50 per cent above rating, but in a number of large installa- 
tions the chimneys have been designed on the coal-burning basis so as 
to provide sufficient capacity in case it proves necessary at a future 
date to revert to the use of coal. 

130. Classification of Chimneys. — Chimneys may be grouped into 
three classes according to the material of construction: 

1. Steel. 

2. Reenforced concrete. 

3. Masonry. 

Steel chimneys have many advantages and are finding much favor 
in large power plants, especially where economy of space warrants 
the erection of the stack over the boiler, in which case the structural 
work of the boiler setting answers for both boiler and chimney. 
Among the advantages are (1) ease and rapidity of construction; 
(2) less weight for a given internal diameter and height; (3) less sur- 
face exposed to the wind; (4) lower cost; (5) smaller space required; 
(6) slightly higher efficiency if properly calked, for there can be no 
infiltration of cold air as is likely through the cracks in masonry. The 
chief disadvantage is the cost of keeping the stack well painted to 
prevent rust and the corrosive action of the sulphur in the coal. 

Steel chimneys may be 

1. Guyed. 

2. Self-sustained. 






CHIMNEYS 



219 



131. Guyed Chimneys. — Guyed sheet-iron or steel chimneys or 
stacks held in position by guy wires are employed in small sizes on 
account of their relative cheapness. They seldom exceed 52 inches in 
diameter and 75 feet in height. A heavy foundation is unnecessary, 
and the stack may be supported by the boiler breeching. The small 
short stacks are ordinarily riveted in the shop, ready for erection, 
larger sizes being shipped in sections and riveted at the place of instal- 
lation. The guy wires are usually fastened to an angle iron or band at 
about two-thirds the height, and anchored at a distance from the 
base equal to the height of the band above the ground. 

For very tall stacks two sets of guys are used, from four to six 
wires being fastened to each band, and designed to withstand a wind 
pressure of 30 pounds per square foot of projected area of the stack. 
Turn-buckles are employed to equalize tautness. Table 28 gives the 
thickness of material, with approximate cost and weight, of guyed 
stacks of different heights and areas. 

132. Self-Sustaining Steel Chimneys. — Steel chimneys over 52 
inches in diameter are usually self-supporting. They may be built 
with or without a brick lining, but the lining is preferred, since it 
prevents radiation and protects the inside from the corrosive action 
of the flue gases. Since the lining plays no part in the strength of the 
chimney, it is made only thick enough to support its own weight, and 
usually of a low-grade fire brick or carefully burned common brick or 
both. In average practice the fire brick extends 20 or 30 feet above 

TABLE 28. 
APPROXIMATE WEIGHT AND COST OF GUYED SHEET-STEEL CHIMNEYS. 



Height, Feet. 


Diameter, Inches. 


Thickness of Shell, 
B.W.G. 


Approximate Weight 
per Foot, Pounds. 


40 


18 


16 


13 


45 


20 


16 


14 


45 


22 


14, 16 


20, 15 


50 


24 


14, 1& 


22, 16 


50 


26 


14 


23.5 


55 


28 


14 


25 


60 


30 


12, 14 


34, 27 


65 


32 


12, 14 


36, 28 


70 


34 


10, 12 


48, 39 


75 


36 


10, 12 


51, 41 



Approximate cost per pound, 3.5 cents to 6.5 cents, including cost of sections 
riveted and punched, ready for assembling, the higher figure referring to the 
smaller stacks. 



220 STEAM POWER PLANT ENGINEERING 

the breeching, the remainder of the lining being of common brick. In 
chimneys up to 80 inches internal diameter, the upper course is 4£ 
inches thick, and increases 4J inches in thickness for each 30 to 40 feet 
to the bottom. In larger chimneys about 8 inches is the minimum 
thickness. The lining is generally set in contact with or close to the 
shell, though a space of 1 to 2 inches is sometimes left between the 
brickwork and the shell to allow for expansion. This space is occa- 
sionally filled with sand. 

Self-sustaining stacks may be straight or tapered, and are generally 
made with a flared or bell-shaped base whose diameter and length are 
1£ to 2 times the internal diameter of the stack. The base is riveted 
to a heavy cast-iron plate bolted to a concrete foundation of sufficient 
mass to insure stability. 

Fig. 117 gives the details of one of the steel chimneys at the power 
house of the South Side Elevated Railroad, Chicago, 111. 

133. Thickness of Plates. — The sheet is thickest at the bottom, 
decreasing toward the top of the stack. The proper thickness for any 
given section may be determined by treating the shaft as a uniformly 
loaded cantilever, the stresses being expressed by the equation 



Ph 

in which 



-'i-'^w-y- 



P = the total wind pressure in pounds. 

h = length of the chimney in inches to the center of wind pressure 
(h = L/2 for a cylindrical chimney). 

S = safe stress. A low value of 6000 pounds per square inch for 
single-riveted joints and 8000 for double-riveted joints is 
recommended, for the reason that a tube of such large 
diameter with thin walls will hardly fail by rupture according 
to the formula, but by flattening and bending. 

- = sectional modulus. 
e 

D x = external diameter of the shell, inches. 

D 2 = internal diameter of the shell, inches. 

For chimneys under 7 feet in diameter and 150 feet in height the 
thickness of plate should not be less than -ft mcn > nor less than J inch 
for larger sizes. 



CHIMNEYS 



221 



i Orf%AT£ 




.&?ICH ■ SACKING 



FLOOft LEV£l 



Fig. 117. Steel Chimney at the Power House of the South Side Elevated Railroad, Chicago. 



222 



STEAM POWER PLANT ENGINEERING 



It is customary to make the courses about 5 feet in height for con- 
venience in erection. 

Table 29 gives the dimensions of self-supporting steel stacks as made 
by the Riter Conley Company of Pittsburg, who use the following 
empirical formula in determining the thickness of the shell 



8 t 



M 



0.8ZV 



(43) 



in which 



S x = stress per lineal inch of section considered, 
M = wind moment in inch-pounds, and 
Dj = diameter of the shaft in inches. 

Allowing 8000 pounds per square inch as the safe stress for single- 
riveted joints and 10,000 for double-riveted joints, the required thick- 
ness is found by dividing S ± by 8000 or 10,000. 

Example : Determine the thickness of plate at a section 150 feet 
from the top of a cylindrical steel stack 12 feet in diameter and 200 
feet high. Horizontal seams to be double riveted. 

The total wind pressure on the section is 

150 X 12 X 25 = 45,000 pounds. 
The moment arm is 

-HP X 12 = 900 inches. 

D x = 144 inches; S = 8000 pounds per square inch. 



TABLE 29. 

STEEL STACKS. — SIZES OF RITER CONLEY COMPANY, PITTSBURG. 



Diameter 
of Flue. 


Total 
Height. 


Total 
Weight. 


How Made. 


Ft. In. 
5 6 

7 

8 6 

10 
12 

11 6 

12 


Ft. 
165 

160 

150 
200 
200 

225 

255 


Lb. 
67,000 

79,000 

94,000 
150,000 
175,000 

232,000 

256,000 


40 ft. of & in., 45 ft. of J in., 50 ft. of & in., 30 ft. of 

1 in. 
30 f t. of & in., 50 ft. of J in., 50 ft. of & in., 30 ft. of 

1 in. 
60 ft. of \ in., 60 ft. of & in., 30 ft. of f in. 
90 ft. of \ in., 60 ft. of & in., 50 ft. of f in. 
35 ft. of I in., 35 ft. of & in., 35 ft. of & in., 35 ft. 

of H in., 35 ft. of f in., 25 ft. of $f in. 
40 ft. of \ in., 40 ft. of & in., 40 ft. of & in., 40 ft. of 

H in., 40 ft. of f in., 25 ft. of ^ in. 
75 ft. of | in., 65 ft. of & in., 55 ft. of f in., 35 ft: 

of ^ in., 25 ft. of | in. 



CHIMNEYS 223 

Substituting these values in equation (42), 

45,000 X 1800 _ 8000 x ZJA /144 4 - DA 



32 V 144 



D a = 143.36. 

Now t = Dl ~ D » 

2 

= 144 - 143.36 

2 

= 0.32 inch. 

The nearest commercial size lies between nine thirty-seconds and five- 
sixteenths. 

The Riter Conley formula gives for this section 

s = M 45,000 X 900 

1 0.8 D* 0.8 X 144 2 

= 2440 pounds. 

, S, 2440 QnK . r 

t = T^h. = ^^ = -305 inch. 
8000 8000 

134. Riveting. — The diameter of rivets should always be greater 
than the thickness of the plate but never less than one-half inch. The 
pitch should be approximately 2\ times the diameter of the rivet, and 
always less than 16 times the thickness of the plate. Single-riveted 
joints are ordinarily used on all sections except the base, where the 
joint should be double riveted with rivets staggered, although in very 
large stacks all horizontal seams are double riveted to give greater 
stiffness to the shaft. 

135. Stability of Steel Chimneys. — The wind being ordinarily the 
only force tending to overturn the stack, and the chimney being rigidly 
bolted to the foundation, a condition of stability requires that 

(W c + Wf) — be equal to or greater than P t — - + h ). (44) 

in which 

W c = weight of the chimney in pounds. 

Wf = weight of the foundation. 

P = total wind pressure in pounds. 

D, H, and h, in feet, as indicated in the figure. 



224 



STEAM POWER PLANT ENGINEERING 



Expressed graphically: Lay off GP, Fig. 118, equal to the total wind 
pressure in direction and amount and acting at the center of pressure 

of the shaft; lay off GW equal to the 
weight of the stack and foundation; 
find the resultant GR and produce it 
to intersect the base line as at R'; if R / 
falls within the inner third of the 
base the stack is stable, provided, of 
course, that the chimney is properly 
designed and constructed. Therefore 
the heavier the combined weight of the 
chimney and its foundation the more 
stable the structure. (See also para- 
graph 140.) 

D in Fig. 118 varies from one-tenth 
to one-fifteenth H, depending upon the 
character of the subsoil. For the 
ordinary concrete foundation, Christie 
("Chimney Design and Theory," p. 57) 
gives as an average value for D 

H 2 d 




D = 



26,000 



+ 10. 



(46) 



Fig. 118. 



Detailed Description of Steel Chimneys : Eng. 

Rec, Feb. 15, 1902, p. 146, July 5, 1902, p. 2, 

April 23, 1904, p. 523, Sept. 10, 1904, p. 314; 

Power, Dec, 1905, p. 716, April, 1905, p. 231, Jan., 1902, p. 6; Engr. U.S., Sept. 1, 

1904, p. 591, June 15, 1905, p. 403; Eng. News, July 20, 1905, p. 64. 



136. Brick Chimneys. — By far the greater number of power-plant 
chimneys are of brick construction and usually of circular section, 
though octagonal, hexagonal, and square sections are quite common. 
The round chimney requires the least weight for stability, and the others 
in the order mentioned. Taking the total wind pressure on the flat 
surface of a square stack as unity, the effective pressure for the same 
projected area will be 0.75 for the hexagonal, 0.6 for the octagonal, and 
0.5 for the round. 

Brick chimneys may be divided into two general classes : 

1. Single shell, Fig. 119, and 

2. Double shell, Fig. 120. 

The double shell is the most common and consists of an outer shaft 
of brickwork and an inner core or lining extending part way or 
throughout the entire length of the shaft. 



CHIMNEYS 



225 





h-*VH 








k*fH 






H 


I 




& 


/6-S" 


&> 






































1 
















e& 


- 




/ff-S' 




; 2 
































































tor-* 


- 




/6*-S m 














3 








i 












































*3'-~ 


^ 




/S'-S* 




4 








i 










































IS'h- 






/G'~S" 
















5 
















■ 


























„- 


«. 




/6'-S" 












s 
s 




6 






k 








F 














TOP OF 
FOUNDATIONAL! 



TOTAL HEIGHT 
ABOVE FOUNDATION 

2.00FF 




SECTION ON A-A SECTION ON BS 

Fig. 119. Custodis Radial Brick Chimney. 



226 



STEAM POWER PLANT ENGINEERING 







BBB 
BOB 



Fig. 119a. Custodis Radial Brick. 



The single shell is the general construction where carefully burned 
and selected brick not easily affected by the heat are used. As the 

inner core or lining is independent 
of the outer shell and has no part 
in the strength of the chimney, the 
rules for determining the thickness 
of the walls are practically the same 
for both single and double shell. 

137. Thickness of Walls. — The 
thickness of the wall should be such 
as to require minimum weight of 
material for the proper degree of 
stability, due consideration being 
paid to the practical requirements of construction. The thickness 
does not vary uniformly, but decreases from bottom to top by a 
series of steps or courses as in Fig. 121. In general, the thickness at 
any section should be such that the resultant stress of wind and weight 
of shaft will not put the masonry in tension on the windward side or 
in excessive compression on the leeward side. 

For circular chimneys using common red brick for the outer shell 
the following approximate method gives results in conformity with 
average practice: 

* = 4 + 0.05 d + 0.0005 H, (47) 

where 

t = thickness in inches of the upper course, neglecting ornamenta- 
tion, and should, of course, be made equal to the nearest 
dimension of the brick in use. Ordinary red bricks measure 
8^x4x2. 

d = clear inside diameter at the top, inches. 

H= height of stack, inches. 

Beginning at the top with this thickness, add one-half brick, or 
4 inches, for each 25 or 30 feet from the top downwards, using a batter 
of 1 in 30 to 1 in 36. 

The minimum value of t for stacks built with inside scaffolding 
should be 7 inches for radial brick and 8^ inches for common brick, 
as a thinner wall will not support the scaffold. Radial brick for chim- 
neys are made in several sizes, so that the thickness of the walls when 
they are used increases by about 2 inches at the offsets. 

For specially molded radial brick or for circular shells reenforced 
as in Fig. 120 the length of the different courses may be much less than 



CHIMNEYS 



227 




Fig. 120. Brick Chimney at the Power Plant of the Armour Institute of Technology. 



228 



STEAM POWER PLANT ENGINEERING 



stated above. The external form of the top is a matter of appearance, 
and may be designed to suit the taste, but should be protected by a cast- 
iron or tile cap and provided with lightning rods. Ladders for reach- 
ing the top of the chimney are generally located inside of brick stacks 
and outside of steel ones. 

Professor Lang's rule (Eng. Rec, July 20, 1901, p. 53) for determining 
the length of the different courses is (Fig. 121) 



- o ( 



2(X+6(H + 0.1056G 



0.453 p - 18 



,), 



+ 2.5 1 + 656 tan a - 0.007 H 



(48) 




in which 



h = length of the course under consideration. 
C = constant = 1 for a circular, 0.97 for an 

octagonal, and 0.83 for a square chimney. 
i = increase in thickness for each succeeding 

section in feet. 
G = weight per cubic foot of brickwork. 
p = wind pressure, pounds per square foot. 
a = angle of the internal batter. 
All other notations as indicated in Fig. 121. 

For chimneys over 100 feet in height he 
recommends that 100 be used instead of the 
actual height, since the critical point will be in 
one of the lower sections and not at the base. 

If a value of h is obtained which is not con- 
tained an even number of times in H, it may be 
slightly increased or decreased so as to effect 
this result. 

To determine the stresses at any section the 
shaft is treated as a cantilever uniformly loaded 
with a maximum wind pressure of 25 pounds per 
square foot. If the tension on the windward 
side subtracted from the compression leaves a positive remainder, the 
chimney will be stable; if the remainder is negative, the masonry will 
be in tension, which it withstands but feebly. The sum of the com- 
pressive stresses on the leeward side due to wind pressure and weight 
must be less than the crushing strength of the masonry. The practice, 
however, of assuming a fixed value for allowable pressure irrespective 
of the height of the stack gives dimensions that are too low for small 



CHIMNEYS 229 

stacks and too high for large stacks. According to Professor Lang, 
compressive stress on the leeward side in pounds per square inch with 
single chimneys should not exceed 

p = 71 + 0.65 L, (49) 

where 

p = pressure in pounds per square inch. 

L = distance in feet from top of chimney to the section in question. 

With double shell p = 85 + 0.65 L. (50) 

The tension on the windward side should not exceed, 

For single shell: p = (18.5 + 0.056 L). (51) 

For double shell: p = (21.3 + 0.056 L). (52) 

Example: Determine the maximum stress in the outer fibers of 
the brickwork at the base of section 8 of the chimney illustrated in 
Fig. 119 when the wind is blowing 100 miles an hour. Assume the 
weight of the brickwork 120 pounds per cubic foot. 

A wind velocity of 100 miles per hour is estimated to exert a pressure 
of 50 pounds per square foot on a flat surface and approximately 25 
pounds per square foot of projected area on a cylindrical surface. The 
height of the chimney to section 8 is 131.4 feet. The projected area 
as computed from the figure is 1800 square feet. Hence p, the total 
wind pressure, is 1800 X 25 = 45,000 pounds. The volume of brick- 
work above section 9 may be calculated, and is 6150 cubic feet, hence 
the weight W = 6150 X 120 = 738,000 pounds. 

The area of the joint at this section is 75.3 square feet, therefore the 
pressure due to the weight of the superimposed brickwork is 738,000 
divided by 75.3 = 9800 pounds per square foot. To find the stress 
due to the wind pressure, substitute the proper values in equation (42) : 

Ph = S I =0.0983 ( Dl *Z D4 ) S. 
Here 

P — 45,000 as computed above. 

h = 55 feet. (Found by laying out the section and locating the 

center of gravity.) 
D x = 16.2. 
D = 12.9. 

Whence 

45,000 X 55 = 0.0983 16 ' 2 * ~ * 2 ' 9 * S. 

16.2 

From which S = 9907 pounds per square foot . 



230 STEAM POWER PLANT ENGINEERING 

The net stress on any part of the section is the resultant of that due 
to the weight of the stack and that caused by the wind, the net stress 
on the windward side being 

9907— 9800 =107 pounds per square foot, 

which is evidently a tensile stress and should never exceed the value 
given by formula (51): 

p = (18.5 + 0.056 L) 

= (18.5 + 0.056 X 131.4) 

= 25.8 pounds per square inch 

= 3715 pounds per square foot. 

The net compressive stress on the leeward side is 9800 + 9907 
= 19,707 pounds per square foot, which should not exceed that given 
by formula (49) : 

p = 71 + 0.65 L 

= 71 + 0.65 X 131.4 

= 156.4 pounds per square inch 

= 22,521 pounds per square foot. 

(See also analysis of steel-concrete chimney, paragraph 142.) 

138. Core and Lining. — The core or lining of a brick chimney is 
commonly carried to the top of the shaft, though it sometimes extends 
only part of the distance. The inside diameter is generally uniform, 
the offsets being made on the outside. The core and outer shell 
should be independent to prevent injury due to expansion of the core. 
The rules for the thickness of lining in steel chimneys apply also to brick 
chimneys. The batters for the inner and outer shells should be such 
as to allow at least 2 inches clearance between the two shafts at the 
top, and the top should be protected by an iron ring or by a projecting 
ledge from the outer shell. 

139. Materials for Brick Chimneys. — Brick for the external shaft 
should be hard burned, of high specific gravity, and laid with lime 
mortar strengthened with cement. Lime mortar itself is more resist- 
ant to heat, but hardens slowly and may cause distortion in newly 
erected stacks, and hence should be used only when a long time is 
taken in building. Mortar of cement and sand alone is not to be 
recommended, since it does not resist heat well and is attacked by 
carbon dioxide, particularly in the presence of moisture. A mortar 
consisting of 1 part by volume of cement, 2 of lime, and 6 of sand 



CHIMNEYS 



231 






may be used for the upper brickwork, 1, 2\, and 8 respectively for the 
lower part, and 1,1, and 4 respectively for the cap. The harder the 
brick the more cement is necessary, as lime does not cling so well to 
hard, smooth surfaces. The inner core may be 
constructed of second-class fire brick, since the 
temperature seldom exceeds 600 degrees F. 
Lime mortar is invariably used for the core. 
140. Stability of Brick Chimneys. — When there 
is no wind blowing and the chimney is built sym- 
metrically about a vertical axis the pressure 
due to weight is uniformly distributed over the 
x bearing surfaces, and the center 



Z 






1 



w 



r *- 

D 



7 A 



(A) 



of pressure lies in the line XX, 
Fig. 122. But when the wind 
blows the pressure exerted tends 
to tilt the shaft as a whole- 
column in the direction of the 
current, and the pressure de- 
creases from the windward to 
the leeward side of the base, 
until, with a sufficiently high 
velocity of wind, it may become 
zero, in which case the center 
of pressure moves a distance q 
FlG - 122 - towards the leeward side of 

the base. As soon as the pressure at A becomes zero the joint begins 
to open (assuming no adhesion between chimney and base) and the 
shaft is evidently in the condition of least stability. The distance q 
through which the center of pressure has moved is called the radius 
of the statical moment. For any column it may be shown that 



(B) 



q = -j- . (Rankine, " Applied Mechanics," p. 229.) (53) 

Ae 

in which 

/ = moment of inertia of the section. 

A = area of the section. 

e = distance from the center of the shaft to the outer edge of the 

joint. 

D 



Thus for circular section, 



For square section } 



a 8 
D 

9 "T 



232 STEAM POWER PLANT ENGINEERING 

D 2 + d 2 



For annular circular ring, q = 

For hollow square, q = 



8D 

D 2 + d 2 

6D 



The relationship between weight of shaft and wind pressure for the 
condition of least stability is 

Ph = Wq, (54) 

in which 

P = total wind pressure, pounds. 

h = distance in feet from the base line of the section under con- 
sideration to center of gravity of that section. 
W = weight of shaft in pounds above the assumed base line. 
q = radius of the statical moment. 

The condition of least stability for round chimneys requires, there- 
fore, that 

D 2 4- rl 2 

Pk = W 8D (55) 

For many purposes it is sufficiently accurate to assume D = d, and 
equation (55) becomes 

Ph = W — for round chimneys. (56) 

Ph = W — for square chimneys. (57) 

o 

The rule commonly used in Germany, and which is finding much 
favor with engineers in the United States, gives for the condition of 
least stability 



w(^R +jA= Ph. (Eng. Rec., July 27, 1901, p. 82.) 



(58) 



Notations as in Fig. 122, all dimensions in feet. 

This permits of a lighter chimney than equation (55), and the maxi- 
mum wind pressure may be assumed to put the joint on the wind- 
ward side in tension or even to permit a slight opening of same. 

A rule of thumb for stability is to make the diameter of the base one- 
tenth of the height for a round chimney; for any other shape to make 
the diameter of the inscribed circle of the base one-tenth of the height. 

The factor of stability is the quotient obtained by dividing the value 
of q from formula (54) by that from (53). If less than unity, the 
chimney is in tension at the outer fiber on the windward side, and must 
be redesigned unless the tension is less than that allowed by equation (51). 
Calculations for stability should be made for various sections. 



CHIMNEYS 233 

Example: Analyze the chimney illustrated in Fig. 119 for stability 
at say section 8, the following data referring to the portion above the 
base line of this section. 

From the drawing : 

Projected area of the stack, 1800 square feet. 

Volume of brickwork, 6150 cubic feet. 

Outside diameter of base, 16.2 feet. 

Inside diameter of base, 12.9 feet. 

Center of pressure to base line, 55 feet. 

Total height above base line, 131.4 feet. 

Maximum total wind pressure : 

P = 1800 X 25 = 45,000 pounds. 

Weight of shaft : 

W = 6150 X 120 = 738,000 pounds. 

For stability, according to equation (55), 

Substituting the proper values : 

Ph = 45,000 X 55_=_2,475,000 foot-pounds. 

W D * + d ' = 738,000 I 16 - 2 ' + J 2 ' 92 ) = 2,441,000. 
8 D ' \ 8X 16.2 / 

While Ph is slightly greater than W — — — — , for practical purposes 

8 D 

the shaft at this section would be called stable under maximum allow- 
able wind pressure. 

For stability, according to equation (58), 
Ph < W(iR + Jr). 
Ph = 2,475,000, as determined above. 
6\45\ 
4 ) 
= 4,177,000. 

Ph is therefore considerably less than W(i R + J r), and the con- 
dition imposed in equation (58) is more than fulfilled. 

Detailed Description of Brick Chimneys : At Paris Exposition, Eng. Rec, Feb. 
17, 1900, p. 155; Boston Elevated Ry., Eng. Rec, Dec. 22, 1900, p. 593; Plymouth 
Cordage Co., Eng. Rec, May 18, 1901, p. 466; Boston, Edison Co., Eng. Rec, Oct. 
10, 1903, p. 438; Smelting Works, Freiburg, Germany, Power, Aug., 1900, p. 13; Ter- 
minal R.R. Assn. of St. Louis, Mo., Cassier's Mag., Jan., 1898, p. 261; Cambridge 
Electric Light Co., Engr. U.S., May 15, 1904, p. 331; Interborough Rapid Transit 
Co., Engr. U.S., Nov. 1, 1904, p. 737; Metropolitan St. Ry. Co., Power, March, 1899, 
p. 1. 



W(i R + J r) = 738,000 ( y + 



234 STEAM POWER PLANT ENGINEERING 

141. Custodis Radial Brick Chimney. — Fig. 119 gives the details of 
a 200 x 10 foot radial brick chimney constructed of special molded 
radial brick, formed to suit the circular and radial lines of each section, 
thus permitting them to be laid with thin, even mortar joints. The 
blocks are much larger than common brick and the number of joints is 
proportionately reduced. They are molded with vertical perforations, 
as shown in Fig. 119a, which permits thorough burning, thereby increas- 
ing the density and strength and at the same time reducing the weight 
of the block. In laying, the mortar is worked into the perforations 
about one-half inch. The first 60 feet above the base are octagonal in 
section, with 36-inch walls, and the balance of circular section, with walls 
tapering gradually from 22 inches to 7 J inches in thickness. A radial 
brick lining extends 60 feet from the base as indicated. The chimney 
was designed to furnish draft for a 3500-horse-power boiler plant and 
cost, erected, $8,800. The entire weight of the chimney exclusive of 
foundation is 870 tons. 

Radial brick chimneys without the inner lining are likely to be 
unduly affected by heat. 

The tallest chimney in the world (1907), located at Great Falls, 
Mont., is of the Custodis type, and is used for leading off the gases from 
the smelter plant of the Boston and Montana Consolidated Copper and 
Silver Mining Company. The height above the top of the foundation is 
506 feet, and the internal diameter at the top 50 feet. The chimney 
and foundation cost approximately $200,000. 

Custodis Chimney Details: Eng. Rec, Oct. 1, 1904, p. 385; Power, May, 1900, 
p. 12. 

142. Steel-Concrete Chimneys. — The use of concrete reenforced with 
iron or steel for the construction of chimneys is rapidly increasing. The 
advantages claimed for this class of stack are: 

1. Light weight of the whole structure, being but one-third as great 
as an equivalent common brick chimney. The space occupied is much 
less than with either brick or steel stack, on account of the thinness of 
walls at the base and the absence of any flare or bell. 

2. Total absence of joints, the entire structure including foundation 
being a monolith. 

3. Great resisting power against tension and compression. 

4. Rapidity of construction. May be erected at an average rate of 
six feet per day. 

5. Adaptability of the material to any form. 

This type of chimney being comparatively new, little data concern- 
ing depreciation are available, but some which have been in use ten 
years show little or no deterioration. 



CHIMNEYS 



235 




GRADE 



Fig. 123. Weber Reenforced Concrete Chimney. 



236 



STEAM POWER PLANT ENGINEERING 



Fig. 123 gives the details of a Weber steel-concrete chimney erected 
at Portland, Oregon, for the Portland General Electric Company. The 
entire structure, foundation and shaft, is a. monolith, 238 feet in total 
height and 12 feet internal diameter, weighing only 889 tons. It 
occupies but 168 square feet of ground space at the grade level. The 
weight not including foundation is 470 tons. The stack was erected 
complete in 58 working days, and cost approximately $13,000. 

The cement used was German Portland mixed with select bank sand 
in proportion of one to three, gravel or crushed stone being used only 
in the foundation below the ground. The mortar was used medium 
dry and tamped in the form around the steel reenforcement. 

The shaft is of the double-shell type, with inner core extending 70 
feet above the grade. The core is but 4 inches in thickness at the 

base, and the outer shell 8 inches. 
Both inner and outer shell are 
reenforced with vertical T bars, 
lJxlJx T \ inch, of low-carbon 
Bessemer steel, spaced at the base 
24 inches between centers in the 
inner core and 4 inches in the 
outer shell, and increasing in 
spacing to the top, where the dis- 
tance between the bars is 12 
inches. The horizontal rings are 
1 x 1 x J T's spaced 18 inches 
between centers in the core and 
36 inches in the outer shell. The 
steel bars vary from 16 to 30 feet 
in length, and where they meet 
lengthwise are lapped not less 
than 24 inches. The use of differ- 
ent lengths of steel prevents the 
laps from concentrating in any 
given section. 

The tallest chimney of this type 

(1907) was erected for the Butte 

Reduction Works at Butte, Mont. 

Its height is 350 feet and inside 

Fig. 124. diameter 18 feet. 

The following strain sheet gives the Weber Company's analysis of 

the chimney illustrated in Fig. 123, and is based on a wind pressure 

of 50 pounds per square foot. Notations as in Fig. 124. 



19.75 Sq.Ft. 




CHIMNEYS 237 

Weights. 

Wf = weight of foundation 



= (k 2 K + l * + l * h\ 150 
= /30 2 - 



2 + 3 ° 2 + 15 ' 3^ 150 



= 523,200 pounds. 
150 = weight per cubic foot of concrete. 
W e = earth weight on foundation 

= {lih 6 — (volume of foundation)} 100 

= (7200 cubic feet - 3995 cubic feet) 100 

= 320,500 pounds. 
100 = weight per cubic foot of earth. 
W = weight of shaft 

= {A t (h A + h 3 ) + A 2 (h, + h a ) + AJi 5 } 150 

= {38.5 (72 +3) + 13 (72 + 3) + 19.75-158} 150 

= 934,950 pounds. 
W t = total weight 

= W f + We + W = 1,778,650 pounds (889 tons). 

Section at Grade G r . 

I. Wi = weight of outer shell and single shell above section 
= (AJit + A 3 hr ) 150 
= (28.5 • 72 + 19.75 • 158) 150 
= 775,806 pounds. 

II. r = radius of statical moment 






14.66 



= 3.35 feet. 



III. P = wind pressure on chimney 

= 14.66 X 72 X 25 + 13 X 158 X 25 
= 77,738 pounds. 



238 STEAM POWER PLANT ENGINEERING 



M = wind moment on section 
5( 
2 



: ^ + (/)Af)^ + |) 



14.66 X 72 X y X y +( 13 x 158 X |°-)( 72 + 



158> 
2 > 

= 8,703,818 foot-pounds. 



IV. N = statical moment 
= rWi. 

= 3.35 X 775,806 
= 2,598,950 foot-pounds. 

V. B = bending moment 
=-- M -N 

= 8,703,818 - 2,598,950 
= 6,104,868 foot-pounds. 

VI. — = section modulus 

e 

= 0.0982 ( D *~ dl *) 

= o 0982 ( U M X 12 * ~ 13 - 33 X 12 * \ 
\ 14.66 / 

= 169,703. 
VII. z = tension per square inch sectional area 

= 12B ^ - 

e 

= 12 X 6,104,868 -^ 169,703 
= 432.5 pounds. 

VIII. Z = total tension 

= 144 A,z 

= 144 X 28.5 X 432.5 
= 1,825,015 pounds. 

IX. s = area steel required 

= — • (a = sectional strain on steel) 
a 

= 16,000 pounds per square inch 

= 114.2 square inches. 



CHIMNEYS 239 

X. K = number of bars 

= — ■ (x = 0.45 square inches = area of one bar) 
x 

= 252 bars. 

For Stability. 

XI. L = length of one side of base. 

= 8,703,818 6 
1,778,650 

= 29.4 feet. 



Section 42' 


0" above Grade. 


I. 


W t 


= 596,250 pounds. 


II. 


r 


= 3.35 feet. 


III. 


P 


= 62,295 pounds. 


IV. 


N 


= 1,997,438 foot-pc 


V. 


B 


= 3,763,889 foot-pc 


VI. 


I 
e 


= 169,703. 


VII. 


z 


= 222 pounds. 


VIII. 


Z 


= 911,088 pounds. 


IX. 


s 


= 57 square inches, 


X. 


K 


= 127 bars. 



M= 5,761,325 foot-pounds. 



Section at Offset. 

I. W t = 468,000 pounds. 
II. r = 3 feet. 

III. P = 51,350 pounds. M= 4,056,650 foot-pounds. 

IV. N = 1,404,000 foot-pounds. 
V. B = 2,652,650 foot-pounds. 

VI. I = 102,041. 

e 

VII. z = 311 pounds. 

VIII. Z = 786,000 pounds. 

IX. s =55.5 square inches. 

X. K = 123 bars. 



240 STEAM POWER PLANT ENGINEERING 

Section 50' 0" from Top. 

I. W t = 148,125 pounds. 
II. r = 3 feet. 

III. P = 16,250 pounds. M= 406,250 foot-pounds. 

IV. N = 444,365 foot-pounds. 

Since the statical moment N is greater than the wind moment M, 
there is no bending moment B, so no steel is required, the chimney 
above this section standing of its own weight. However, thirty-two 
bars are continued to the top. 

Detailed Description of Concrete-Steel Chimneys : Weber Chimney, Power, July, 
1907, p. 476; Burt Portland Cement Co., Eng. News, Dec. 29, 1904, p. 579; Leiter 
Coal Mines, Eng. Rec, March 21, 1904, p. 661, Power, Dec, 1904, p. 787; Pacific 
Electric Ry. Co., Los Angeles, Cal., Eng. News, April 2, 1903, p. 308; Laclede Fire 
Brick Co., Eng. News, April 2, 1903, p. 310; Tall Concrete Chimneys, Eng. News, 
July 20, 1905, p. 57, Aug. 3, 1905, p. 120, Feb. 15, 1906, p. 165, Oct. 11, 1906, 
p. 387; Butte Reduction Works, Eng. Rec, Feb. 3, 1906, p. 124; New Types, Eng. 
Rec, Dec. 15, 1906, p. 670. 

143. Breeching. — The area of the flue or breeching leading from the 
boilers to the chimney is generally made equal to or a little larger than 
the internal area of the chimney, 20 per cent greater being an average 
figure. The flue may be carried over the boilers or back of the setting, 
or even under the fire-room floor, but in any case should be as short as 
possible and free from abrupt turns. Short right-angled turns reduce 
the draft approximately 0.05 inch for each turn, and a convenient rule 
is to allow 0.1 inch loss for each 100 feet of flue if of circular cross 
section and constructed of steel, and double this amount for brick flues 
of square section. The cross section of the flue need not be the same 
throughout its entire length, but may be tapered and proportioned to 
the number of boilers. Where two flues enter the stack on opposite 
sides, a diaphragm is inserted as indicated in Fig. 119. Flues should 
be covered with heat-insulating material. 

144. Chimney Foundations. — On account of the concentration of 
weight on a small area the foundation of a chimney should be carefully 
designed. In most cities the building laws limit the maximum loads 
allowed for various soils and materials, and although they vary con- 
siderably the average is approximately as follows: 

Material. Safe Load, Lb. per Sq. Ft. 

Hard-burned brick masonry, cement mortar, 1 to 2 20,000-30,000 

Hard-burned brick masonry, cement mortar, 1 to 4 18,000-24,000 

Hard-burned brick masonry, lime mortar 10,000-16,000 

Concrete, 1 to 8 8,000-10,000 



CHIMNEYS 241 

Kind of Soil. Safe Load, Tons per Sq. Ft. 

Quicksands and marshy soils 0.5 

Soft wet clay 1.0 

Clay and sand 15 feet or more in thickness 1.5 

Pure clay 15 feet or more in thickness 2.0 

Pure dry sand 15 feet or more in thickness 2.0 

Firm dry loam or clay 3.0- 4.0 

Gravel well packed and confined 6.0- 8.0 

Rock broken but well compacted 10.0-15.0 

Solid bed rock Up to -g- of its ultimate crushing strength. 

Tons per Pile. 

Piles in made ground 2.0 

Piles driven to rock or hardpan 25.0 

Chimney foundations as a rule are constructed of concrete except 
where the low sustaining nature of the soil necessitates the use of piles 
or a grillage of timber or steel. For masonry chimneys the foundation 
is designed to give the necessary support to the shaft without particular 
reference to its mass or distribution, as the shape of the foundation has 
virtually no effect on its stability as a column. In steel and reenforced 
concrete chimneys the shape and weight of the foundation are a function 
of the desired factor of stability, since the shaft is securely anchored 
to the foundation and the two form practically one mass. The founda- 
tion should be designed to fulfill the conditions in formula (46) in 
addition to the requirements for mere support. 

Table 30 gives the least diameter and depth of foundation for steel 
chimneys of various diameters and heights. 

145. Chimney Efficiencies. — The chimney as a mover of air has a 
very low thermodynamic efficiency. Compared with that of a fan its 
performance is very poor, and mechanical-draft concerns sometimes 
use this as an argument. 

Example: A chimney 200 feet high and 10 feet in diameter furnishes 
draft for a battery of boilers rated at 3500 horse power. Average 
outside temperature 60 degrees F. ; temperature of flue gases 500 degrees 
F.; calorific value of the fuel 14,000 B.T.U. per pound. Compare the 
thermal efficiency of the chimney as a mover of air with that of a forced 
draft apparatus of equivalent capacity. 

From Table 24 we find that a chimney 200 feet high, with tempera- 
tures as stated above, will furnish a theoretical draft of 1.27 inches, 
equivalent to a pressure of 6.6 pounds per square foot. Neglecting 
friction the height H of a column of external air which would produce 
this pressure is 

H=(^- r ^h, (59) 



242 STEAM POWER PLANT ENGINEERING 

in which 

h = height of the chimney in feet. 

d = density of the hot gases in the stack. 

d 1 = density of the outside air. 

Substitute in (59), 

d, = 0.0763, d = 0.0435, and h = 200. 

TABLE 30. 

SIZES OF FOUNDATION FOR STEEL CHIMNEYS. 



Diameter, Feet. 


Height, Feet. 


Least Diameter of 
Foundation. 


Least Depth of 
Foundation. 


3 


100 


15' 9" 


6' 0" 


4 


100 


16' 4" 


6' 0" 


4 


125 


18' 5" 


7' 0" 


5 


150 


20' 4* 


9' 0" 


5 


200 


23' 8" 


10' 0" 


6 


150 


21' 10" 


8' 0" 


6 


200 


25' 0" 


10' 0" 


7 


150 


22' r 


9' 0" 


7 


250 


29' 8" 


12' 0" 


9 


150 


23' 8" 


10' 0" 


9 


275 


33' 6" 


12' 0" 


11 


250 


24' 8" 


10' 0" 


11 


350 


36' 0" 


14' 0" 



H = 



* 



0763 



0.0763 / 



85.9 feet. 



The theoretical velocity of the air entering the base of the chimney 
under this head is 

v = V2gh 

= V2 X 32.2 X 85.9 
= 74.5 feet per second. 

The weight of the gas escaping per second 

= 74.5 X area of the stack X 0.0763 
= 446 pounds. 

The displacement of this volume of gas is the result of heating it from 
60 to 500 degrees F. Taking the specific heat of the gas as 0.2375, the 
heat necessary to displace 456 pounds per second is 

Heat required = 446 X 0.2375 X (500 - 60) 
= 46,500 B.T.U. per second 






CHIMNEYS 



243 



The work actually performed is that of overcoming a total resistance 
of 6.6 X 78.5 = 518 pounds (78.5 = internal area of the chimney) 
through a space of 74.5 feet; i.e., 

Work done = 74.5 X 518 = 38,591 foot-pounds per second, 
= 49.7 B.T.U. per second. 

Efficiency = 4 = .00107, or about T V of 1 per cent. 

If a fan be substituted for the chimney and we allow say 8 per cent 
for the efficiency of engine and boiler, 40 per cent for the fan, and 25 per 
cent for friction, the combined efficiency will be 

0.08 X 0.40 X 0.75 = 0.024, or 2.4 per cent. 

024 

The fan then will be — '■ = 22.4 times more efficient than the 

0.00107 

chimney as a mover of air. 
146. Cost of Chimneys. — Christie ("Chimney Design and Theory") 

gives the following costs of chimneys 150 feet high and 8 feet internal 

diameter : 

Common red brick approximate cost $8,500.00 

Radial brick do do 6,800.00 

Steel, self-supporting, full lined do do 8,300.00 

Steel, self-supporting, half lined do do 7,800.00 

Steel, self-supporting, unlined do do 5,820.00 

Steel, guyed do do 4,000.00 

The following approximate costs of various sizes of a well-known 
radial brick chimney give an idea of the variation in cost due to in- 
crease in diameter and height : 



Size of Chimney. 




Size of Chimney. 








Cost. 






Cost. 


Height. 


Diameter. 


Height. 


Diameter. 


Feet. 


Feet. 




Feet. 


Feet. 




75 


4 


$1,350.00 


175 


8 


$7,050.00 


75 


6 


1,950.00 


175 


10 


7,925.00 


75 


8 


2,650.00 


175 


12 


8,950.00 


75 


10 


3,725.00 


175 


14 


9,725.00 


125 


6 


3,500.00 


200 


8 


9,250.00 


125 


8 


4,250.00 


200 


10 


10,500.00 


125 


10 


4,675.00 


200 


12 


11,100.00 


125 


12 


5,125.00 


200 


14 


12,500.00 


150 


8 


6,150.00 


250 


10 


16,500.00 


150 


10 


7,125.00 


250 


12 


18,250.00 


150 


12 


7,750.00 


250 


14 


21,500.00 


150 


14 


8,275.00 


250 


16 


24,250.00 



244 



STEAM POWER PLANT ENGINEERING 



TABLE 31. 

PROPORTIONS OF CHIMNEYS FOR FACTORY STEAM BOILERS, COLLECTED 
FROM PRACTICE. (Hutton.) 



Height of 


Internal Dimensions. 


Ratio of 


Thickness of Walls. 


Chimney 

above the 

Ground in 

Feet. 








Bottom to 

Top. 
Internal 

Area. 






Size of Base at the 
Ground Line. 


Siz* 


. of Top. 


Thickness 

at Base in 

Inches at 

Ground 


Thickness 
at the Top 
in Inches. 












Line. 




40 


2' 6" 


1' 


9" sq. 


2.04 


18 


9 


60 


2' 11" 


2' 


0"sq. 


2.12 


18 


9 


70 


3' 4" 


2' 


3"sq. 


2.13 


23 


9 


80 


3' 8" 


2' 


6" sq. 


2.18 


28 


9 


90 


4' 0* 


2' 


9" sq. 


2.27 


28 


9 


100 


4' 8" 


3' 


0" diam. 


2.40 


28 


9 


110 


4' 10" 


3' 


3" diam. 


2.33 


28 


9 


120 


5' 6" 


3' 


6" diam. 


2.40 


28 


9 


135 


6' 0" 


4' 


0" diam. 


2.30 


28 


9 


150 


4' 6" 


3' 


0" diam. 


2.25 


28 


14 


155 


6' 0* 


4' 


6" diam. 


1.78 


56 


14 


160 


9' 0" 


5' 


0" sq. 


3.24 


36 


14 


170 


V 6" 


5' 


0" diam. 


2.25 


36 


14 


180 


6' 4" 


4' 


6" diam. 


2.00 


54 


14 


200 


5' 3" 


3' 


6" diam. 


2.28 


36 


14 


225 


16' 0" 


6' 


6"sq. 


4.00 


36 


14 


250 


19' 0" 


13' 


0" diam. 


2.13 


40 


14 


300 


14' 0" 


9' 


0" diam. 


2.42 


48 


14 


450 


21' 6* 


10' 


2" diam. 


4.35 


59 


14 






CHAPTER VIII. 

MECHANICAL DRAFT. 

147. General. — The intensity of natural draft in a chimney depends 
mainly upon the height of the stack and the temperature of the chim- 
ney gases, and the chimney should be designed to meet the maximum 
requirements, permitting the damper to be partly shut at times. There 
is usually no practicable means of increasing the draft after the maxi- 
mum has been reached. Again, chimney draft is peculiarly susceptible 
to atmospheric influence and may be seriously impaired bv adverse 
winds and air currents. Notwithstanding these apparent limitations, 
by far the greater number of steam power plants depend upon chim- 
neys for draft, and for obvious reasons as will be discussed later. In 
many cases artificial draft has a great advantage and under certain 
conditions is indispensable; it is very flexible and readily adjusted 
to effect various rates of combustion, irrespective of climatic influ- 
ences, and permits any degree of overload without undue expenditure 
of energy. 

Artificial draft may be broadly classified under two heads 

1. The vacuum or induced draft and 

2. The plenum or forced-draft method. 

In the former a partial vacuum is produced above the fire by suitable 
apparatus, and the effect is substantially that of natural draft. 

In the forced-draft system pressure is produced in the ash pit, the 
air being forced through the grate. 

In both systems the artificial draft is usually produced by either 

1. Steam jets or 

2. Centrifugal fans or exhausters. 

148. Steam Jets. — Fig. 125 shows an application of a ring jet to 
the base of a stack. The apparatus is very simple, inexpensive, and 
easily applied. It consists essentially of a ring or a series of concen- 
tric rings of 1-inch or lj-inch pipe, perforated on the upper side with 
T V or J- inch holes, and placed in the base of the stack, so that the 
jets are discharged upward, thus creating a draft independent of the 
temperature of the flue gases. The steam connection to the jet is 
generally made direct to the boiler and not to the steam main, though 
the jet is often produced by exhaust steam. 

245 



246 



STEAM POWER PLANT ENGINEERING 



Fig. 126 illustrates a Bloomsburg jet, which involves to some extent 
the principle of the ejector. 

The increase in draft produced by these devices as ordinarily installed 






Fig. 125. Ring Steam Jet. 



Fig. 126. Bloomsburg Jet. 



is not great, although in locomotive practice where the entire exhaust 
is discharged up the stack an intense draft is obtained. 

Fig. 127 shows the application of a " McClaves argand blower." 




Fig. 127. McClaves Argand Blower. 

The steam is discharged below the grate through a perforated hollow 
ring, as indicated, drawing the air through the funnel by inspiration. 
This creates a powerful draft by forming an air pressure in the ash pit, 



MECHANICAL DRAFT 



247 



and is an especially useful system of forcing fires for boilers which need 
forcing for short periods only. 

Steam jets are very uneconomical, since a large amount of steam is 
required to produce good results. Table 32, based on experiments at 
the New York Navy Yard to determine the best form of steam jet 
for producing draft in launch boilers, shows steam consumptions of 
8.3 to 21.2 per cent of the total steam made. Table 33 gives the steam 
consumption of a number of types of steam jet blowers as determined 
by A. J. Whitham. The best performance is 4.6 per cent and the 
poorest 11.1 per cent of the total boiler steam generated. Steam jets 
below the grate are said to prevent clinker from forming where fine 
anthracite coals are used, and thus to assist in keeping the fire free 
and open. 

Steam jets arranged above the grate and discharging either from the 
side walls, front wall, or bridge wall, oftentimes assist complete com- 
bustion by stirring up the volatile gases and air and insuring a 
thorough mixture, thus affording one of the simplest and frequently a 
very efficient means of furthering smokeless combustion. The action 
is of course purely mechanical, the steam in itself not being a sup- 
porter of combustion; hence if the air supply is deficient the steam jet 
is of no avail unless arranged to carry sufficient air along with the 
steam. 

Fig. 128 illustrates such an application to a hollow bridge wall. 
The top of the wall is fitted with a small cast-iron column M, partially 





GRATE 



Fig. 128. Application of Steam Jets to Hollow Bridge Wall. 



imbedded in the brickwork. A series of 1-inch holes " 00" drilled 
near the top of the casting, furnish exits for the steam and air. A 
steam jet in one end of the column induces air into the iron chamber 
and forces it across the fire in fine streams. Excessive air dilution is 
avoided by partially closing the ash-pit doors and by regulating the 
intensity of the jets. An installation of this type is especially effective 
in connection with coal having a tendency to fuse and seal the air pas- 



248 



STEAM POWER PLANT ENGINEERING 



sages in the grate. Two Stirling boilers at the Armour Institute of 
Technology equipped with this device gave practically smokeless 
combustion at all normal loads, though at heavy overloads it was 
sometimes necessary to slightly open the fire doors. Without the use 
of the jets smoke could not be prevented even at light loads. Analysis 
of the flue gases showed but a slight decrease in the percentage of C0 2 . 

TABLE 32. 
RESULTS OF EXPERIMENTS UPON STEAM JETS AT NEW YORK NAVY YARD.* 





Pounds of Water Evaporated per Hour. 


Index of Jet. 


A 


B 


C 


D 


E 


In boiler making steam 

In boiler supplying jets 

Per cent of steam used 
bv iet 


463.8 
97.5 

21.2 


580.0 
120 

20.7 


361.25 
30 

8.3 


528.5 
63.2 

12.0 


545.00 
76.25 

19.0 







* Annual Report of the Chief of the Bureau of Steam Engineering, U. S. Navy, 1890. 

TABLE 33. 
CONSUMPTION OF STEAM BLASTS COMPARED, t 



Coal. 


Name of Blower. 


Per Cent of Air 

Openings in 

Grate. 


Pounds of Dry 

Coal burned per 

Hour per Square 

Foot of Grate. 


Per Cent of Total 
Steam Generated 

in the Boilers 
that is required 

to operate the 
Steam Blasts. 


Rice 


Young 


11 

11 

7 

11 

11 

26 

11 

11 

7 

7 


25.8 
17.9 
27.0 
27.3 
16.7 
31.4 
16.4 
26.1 
32.5 
45.4 


11.1 


Do 


...do 


7.0 


Do 


Wilkinson 

Young 

....do 


10.8 


Buckwheat 

Do 


10.8 
4.6 


Do 


...do 


8.9 


Do 


McClave 

....do . 


6.7 


Do 


9.3 


Do 


Wilkinson 

....do 


7.8 


Do 


10.2 









t Trans. A.S.M.E., Vol. XVII. — Whitham. 



149. Parson Smokeless Furnace. — The Parson forced-draft system 
for smokeless combustion, applied to a return tubular boiler as illus- 
trated in Fig. 129, comprises a specially designed grate G, depending 
upon a steam jet blower A for draft. Part of the steam is admitted 



MECHANICAL DRAFT 



249 



below the grate and part over the fire through the hollow bridge wall H. 
The supply of air above the grate is regulated by means of damper F. 
The steam to blower A is automatically adjusted by regulator N, 
which is actuated by the steam pressure. The steam to the jet is 
superheated by passing the supply pipe through the setting as indi- 




Fig. 129. Parson's Smokeless Furnace: 



cated. The bridge wall H is provided with an extension platform M 
for holding the unburned fuel when cleaning the fire. 

150. Heinrich Smokeless Furnace. — Figs. 130 and 131 show the appli- 
cation of the Heinrich system of forced draft to a return tubular boiler. 
Hot air is taken from the boiler room above the boilers by a steam jet 
blower at A and forced into the superheating chamber below the 
combustion chamber. From this chamber part of the air is drawn by 
the auxiliary blowers C and forced through tuyeres above the grate, 
the rest passing through an opening beneath the bridge wall into the 
ash pit and up through the bed of fuel. Steam for the blower A and 
the auxiliaries C is supplied through an automatic regulator R, which 
opens when the steam pressure falls below the required value. The 
manufacturers (Heinrich Manufacturing Company, Milwaukee, Wis.) 
sell this apparatus with a guarantee of 15 per cent saving in fuel over 
natural draft, common grate bars, and hand firing. 

151. Fan Draft. — Fig. 132 shows a typical installation of a centrif- 
ugal fan on the forced-draft or plenum principle, the fan creating a 
pressure in the ash pit and forcing air through the fuel. The most 
approved method is to pass the air through the bridge wall, thence 
toward the front of the grate, though it may enter through an under- 
ground duct or through the side of the setting. Forced draft is usually 
adopted in old plants where increased demands for power require that 
the boilers be forced far above their rating to save the heavy expense 



250 



STEAM POWER PLANT ENGINEERING 



of new boilers, or in plants burning refuse, anthracite culm or screen- 
ings, which require an intense draft for efficient combustion. Forced 
draft is also well adapted for underfeed stokers of the retort type, 
hollow blast grates, and the closed fire hole system. The air supply 
may be taken from an air chamber built around the breeching, thereby 
supplying the heated air to the fan and effecting a lower temperature 
in the breeching and a higher temperature in the furnace. The 
objection is sometimes raised against forced draft that the gases tend 




Fig. 130. Heinrich Smokeless Furnace (Sectional Elevation). 




Fig. 131. Heinrich Smokeless Furnace (Sectional Plan). 

to pass outward through the fire door when the fire is cleaned or re- 
plenished, since the pressure in the furnace is greater than atmospheric. 
This objection may usually be overcome by suitable dampers in the 
blast pipe which are closed on opening the fire doors. With a boiler 
plant of 1000 horse power or more the cost of a forced-draft fan, 
engine, and stack will approximate 20 to 30 per cent of the outlay of 
an equivalent brick chimney. The power consumption will depend 



MECHANICAL DRAFT 



251 



upon the character and efficiency of the motor or engine and will range 
from 1 to 5 per cent of the total capacity. 

Induced draft as illustrated in Fig. 133 is perhaps the most com- 
mon substitute for natural draft and is extensively used in street rail- 
way and lighting plants which have high peak loads, being ordinarily 
installed in connection with fuel economizers. The suction side of the 
fan is connected with the uptake or breeching of the boiler or bat- 
teries of boilers and the products of combustion usually exhausted 
through a stub stack. The illustration shows a typical installation in 
which two fans of the duplex type are placed above the boiler setting. 
The fan ducts are generally designed with a by-pass direct to the stack 
to be used in case of accident or when mechanical draft is not required. 




Fig. 132. Typical Forced-Draft System. 

Since the fan handles hot gases it must, under the ordinary con- 
ditions of practice, have a capacity approximately double that of a 
forced-draft fan delivering cold air, but the gases being of lower 
density the power required per cubic foot moved is less. 

With forced draft about 300 cubic feet of air are required per pound 
of coal; with induced draft the fan must handle twice this volume if 
the gases are exhausted at 500 degrees F. or 450 cubic feet if exhausted 
at 300 degrees F., a temperature to be expected in connection with 
economizers, 

The advantages of induced draft over forced draft are very pro- 
nounced. The pressure in the furnace is less than atmospheric, there- 
fore it is not necessary to shut off the draft in cleaning fires or ash pit, 
and the fire burns more evenly over the entire grate area and requires 
less attention than with forced draft. An induced-draft plant costs 
considerably more than forced draft on account of the larger fan 



252 



STEAM POWER PLANT ENGINEERING 



required, but the operating expenses are but little greater. With a 
boiler plant of 1000 horse power or more the cost of a single induced- 
draft fan, engine, stack, etc., will approximate 40 to 50 per cent of the 
outlay required for a brick chimney of equivalent capacity, and the 
double-fan outfit will approximate 50 to 60 per cent. The double-fan 




Fig. 133. Typical Induced-Draft System. 



system is particularly adapted to plants which operate continuously 
and where even a temporary break-down is a serious inconvenience. 

Advantages of Mechanical Draft: Am. Elecn., June, 1898, p. 244, Feb., 1902, 
p. 63; Eng. Rev., Sept., 1901, p. 4; Eng. Mag., April, 1901, p. 81, March, 1900, 
p. 931; Elec. Rev., Lond., Feb. 3, 1899, p. 186; Cassier's Mag., Nov., 1898, p. 48, 
Jan., 1905, p. 252, March, 1906; St. Ry. Rev., July 15, 1901, p. 415; Engr. U.S., 
July 16, 1906, p. 475; Elec. Eng., Aug. 11, 1905, p. 193. 

Application of Mechanical Draft to Stationary Boilers : Power, Dec, 1900, p. 30; 
West. Elecn., Feb. 16, 1901, p. 118; Jour. West. Soc. Engrs., March 19, 1902, p. 271; 
St. Ry. Rev., July, 1899, p. 463; Cassier's Mag., Nov., 1898, p. 48; Elec. Rev., July 
27, 1898, p. 52; Engr. U.S., Jan. 1, 1907. 

152. Performance of Fans. — The first satisfactory theory of centrif- 
ugal fans was promulgated by Daniel Murgue in 1872. He proved 



MECHANICAL DRAFT 



253 



that theoretically the maximum pressure created by a perfect fan is 
equivalent to twice the head which would produce a velocity equal to 
that of the periphery. Thus 



tf =^, 



(60) 



in which 



H = maximum difference in pressure in feet of air, 
u = peripheral velocity in feet per second, and 
g = acceleration of gravity 32.2. 

A and B, Fig. 134, represent Pitot tubes inserted in the discharge 
pipe of a centrifugal blower, A being bent to face the current, while B 
is at right angles to it. A receives the full impact of the stream, and 



H 

JL 



/^\ 



cf 



(A) 






(B) 



ORIFICE CLOSED 



Fig. 134. 



the manometer indicates the total pressure, static and velocity, while 

B registers the static pressure only. With the discharge orifice closed, 

as in Fig. 134", the velocity becomes zero, and the water depression in 

both manometers will be the same, due to the static pressure, which, 

according to Murgue's theory, will be a maximum and, ignoring fric- 

u 2 
tion or eddy currents = — • 

Example : Determine the maximum pressure, in inches of water, 
which a perfect fan would exert with discharge orifice closed; diameter 
of fan 6 feet; r.p.m. 318. 

The peripheral velocity is 

u = 2 nrn = 6.28 X 3 X 318 = 6000 feet per minute. 

= 100 feet per second. 

Substituting in Murgue's formula, 

u = 100 and g = 32.2, 

H = 155! = 310 feet, 
32.2 



254 



STEAM POWER PLANT ENGINEERING 



i.e., the pressure created by the fan would be equivalent to the weight 
of a column of air 310 feet high, or, assuming an air temperature of 
75 degrees F., an equivalent head in inches of water of 
310X0.074495 



144 X 0.0361 



= 4.45 inches. 



(0.074495 = density of air at 75 degrees F. and 0.0361 = pressure pro- 
duced by one inch of water in pounds per square inch.) 

If the discharge orifice be opened to its maximum (Fig. 135) the static 
pressure indicated by manometer B becomes zero, since there is no 

/rS\ n /^n\ n 



%# 



a 



(A) 



%# 



(B) 



ORIFICE WIDE OPEN 



Fig. 135. 

resistance due to the air flow, while the water in A stands at a 
height H the exact equivalent of the velocity head in accordance with 
the hydraulic formula, 

v = V2gH, 

in which v is the velocity of the air in feet per second. 

If the orifice be partially closed, say 50 per cent, as in Fig. 136, B indi- 
cates the static pressure, while A gives the dynamic or total pressure 
due to both velocity and resistance. The difference between A and B 
is therefore the pressure due to velocity alone. By connecting the 
two manometers as indicated in Fig. 136 C the velocity pressure is 
given directly. 

Pressure. — According to Murgue's theory the maximum pressure 
which may be developed by a blower or exhauster varies with the square 
of the speed and may be expressed 

Cdu 2 



V = 



9 



in which 

p = pressure, pounds per square foot. 
d = density of the air, pounds per cubic foot. 
u = peripheral velocity, feet per second. 
C = a coefficient obtained by experiment. 



MECHANICAL DRAFT 



255 



Tables 34 and 35 give the relationship between pressure and speed 
for various sizes of forced and induced-draft fans. 

Fig. 139 shows the relationship between pressure and speed in a 
45-inch Buffalo blower as tested at the Armour Institute of Technology. 

Velocity of Discharge. — The maximum velocity of the air leav- 
ing the tips of the blades varies directly as the peripheral speed, 



V = Ku, 



(63) 



in which 



V = velocity of the air discharged, feet per second. 
K = a coefficient obtained by experiment. 
u = peripheral velocity, feet per second. 



^ 



a 



A) 



/~\ 



/^\ /?n\ 



%& 



(B) 



o 



^ 



(O 



ORIFICE PARTLY CLOSED 



Fig. 136. 



For practical purposes the velocity of discharge with outlet wide open 
may be assumed to be that of the periphery. 

Capacity. — The relationship between capacity and speed, capacity 
and discharge opening for a 45-inch pressure blower is given in Figs. 139 
and 140. 

As will be noted, the capacity varies almost directly with the speed of 
the wheel and the area of discharge as expressed by the equation 



in which 

Q 

B 
A 
D 

N 



Q = BttADN, 

cubic feet discharge per minute, 
coefficient determined from experiment, 
area discharge opening, square feet, 
diameter of the wheel, 
r.p.m. of the wheel. 



(63a) 



256 STEAM POWER PLANT ENGINEERING 

Power. — The power required to drive a fan is proportional to the 
cube of the speed, 

Horse Power = XAN 3 , (64) 

in which 

X = a coefficient determined by experiment. 
A = area discharge outlet, square feet. 
N = r.p.m. 

The marked increase in power required for even a moderate increase 
in speed should be borne in mind in selecting a fan. It is as a rule 
more economical to err in selecting too large a fan than one which must 
be forced above its rated capacity. 

In practice the size of fan is proportioned upon experience rather 
than theory, the usual procedure necessitating the use of curves based 
upon the performance of fans of the type under consideration. 

The curves in Fig. 137 were computed by Mr. F. R. Still of the 
American Blower Company, and give the performance of steel-plate 
fans as manufactured by this company. These curves apply to this 
type and make of fan only, though the difference is not very great 
for any type of centrifugal fan. The " ratio of opening " refers to the 
actual percentage of opening compared with the total discharge. The 
" ratio of effect " is the relative effect produced by restricting the 
discharge. The abbreviations are as follows : 

D.P. = dynamic or total pressure. 
P.V.P. = pressure created by a column of air moving at the 
same velocity as the periphery. 
S.P. = static pressure. 
V.P. = velocity pressure = D.P — S.P. 

Suppose a fan with an unrestricted inlet and outlet delivers 25,000 
cubic feet of air per minute against a head (D.P.) of 0.33 inch with a 
peripheral velocity requiring 6.16 horse power. It appears from the 
curves that if the discharge outlet is restricted to 50 per cent of the 
full area, only 12,500 cubic feet will be delivered; the pressure will be 
increased to 1.03 inches, and the power required drops to 4.84 horse 
power. If the outlet be still further reduced to 20 per cent of the full 
opening the capacity will drop to 5000 cubic feet, the pressure will 
increase to 1.15 inches, and the power will be decreased to 3.45 horse 
power. With a discharge area of 60 per cent, the mechanical efficiency 
is a maximum, and equal to about 43 per cent. With orifice closed 
the horse power required to drive the fan is about 37 per cent of that 
required when discharging the maximum volume of air. 



MECHANICAL DRAFT 



257 



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10 20 30 40 50 60 70 80 90 100 110 120 130 140 
Ratio of Effect Per Cent 

Fig. 137. Performance of Steel Plate Fans. 



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45 In 
Buffalo Blower 
Speed Constant 
1500 R.P.M. 
Discharge Area Variable 


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eet 









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Fig. 138. 



258 



STEAM POWER PLANT ENGINEERING 



Curve " K " in Fig. 137 was determined from the empirical formula 
(based upon Murgue's theorem) 



A-KQ 

in which 

A = area of the inlet orifice, square feet. 

Q = volume of gas, thousands of cubic feet per minute. 

P = draft at the inlet in inches of water. 

K = constant determined by experiment. 



500 



600 



700 



800 900 1000 1100 
.Revolutions per Minute 



Fig. 139. 



(65) 







-2.0 












45 In. 










































Buffalo Blower 

Discharge Area Constant 

Speed Variable 
































-1.8 
-1.7 
-1.6 
-1.5 






























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-50- 
-45- 


6 

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ress 


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1200 1300 1400 1500 



The curves in Figs. 138 and 139 are plotted from tests made at the 
Armour Institute of Technology on a 45-inch Buffalo pressure blower, 
and are characteristic of this type of fan. 

Theory of Fans : Power, May, 1907, p. 287; Engr., Oct. 9, 1903, p. 512; Mach., 
Aug., 1898; Sib. Jour, of Eng., Nov., 1902 ; Heat and Vent., Jan. 15, 1897, July, 1899 ; 
Prac. Engr., Jan. 16, 1903; Mech. Engr., April 18, 1903; Eng. Rec, Oct. 11, 1902. 

Pressure Fans vs. Exhaust Fans: Bulletin Ana. Inst. Min. Engrs., Feb., 1909. 

153. Determination of Size of Fan. — The following analysis, based 
upon a paper on Mechanical Draft by F. R. Still of the American 



MECHANICAL DRAFT 259 

Blower Company, gives a good idea of the usual procedure in deter- 
mining the size of fan for an induced draft installation. (Jour. West. 
Soc. Engr., May, 1902.) 

Example : Determine the size of induced fan and the approximate 
power required to drive it, for a boiler plant rated at 1000 horse 
power; temperature of flue gases 500 degrees F.; heat value of coal 
14,000 B.T.U. per pound; ash 5 per cent; draft required, 1 inch of 
water pressure. 

Assuming a boiler efficiency of 70 per cent, the evaporation will be 
14 I 000 X Q 70 = 1(U5 pounds of water from and at 212 degrees F. 

per pound of coal. 

Since one boiler horse power is equivalent to the evaporation of 
34.5 pounds of water per hour from and at 212 degrees F., the, evapora- 
tion per hour will be 34.5 X 1000 = 34,500 pounds, and the coal 
burned per hour, 

-2^522 = 3400 pounds. 
10.15 

Allowing 18 pounds of flue gas per pound of combustible, 5 per cent 
for ash and 5 per cent for leaks, the fan will have to handle, at 500 
degrees F., approximately 20 X 3400 = 68,000 pounds of gas per 
hour, or 26,000 cubic feet per minute. It is customary, when little is 
known about a plant in which a fan is to be installed, to assume that 
the resistance is equivalent to restricting the discharge outlet 25 per 
cent. Hence in this problem the various factors are referred to a 
" ratio of opening " of 75 per cent (see Fig. 137). 

From formula 65, the area of the inlet should be 

A ,gg, 0-«5X26 , 126 g fe 

Vp l 

which corresponds to a diameter of 48 inches. (K = 0.485 is taken 
from the curves in Fig. 127.) 

The area of the inlet should not exceed 40 per cent of the area of the 
side of the wheel; the latter, then, will be 

— -j- =31.5 square feet, 

which corresponds to a diameter of 76 inches (6.3 feet). 

Referring to Fig. 137, the ratio of dynamic pressure to peripheral 
velocity pressure (D.P. to P.V.P. at 75 per cent opening) is 0.73. The 

peripheral velocity pressure will be — - = 1.37 inches of water. 



260 STEAM POWER PLANT ENGINEERING 

The peripheral velocity is 



U = V2 gH' = 8.03 V#', where H' is the peripheral velocity 
pressure expressed in feet of gas, or, 

x> X 62 5 

Since W = _£. ! — , where p = inches water, 

0.0478 X 12' F 

U = 87.5 Vl.37. 

= 102.5 feet per second. 

= 6150 feet per minute. 

The maximum effective discharge area which an inclosed fan of this 
type may have, and still maintain the pressure equivalent of the per- 
ipheral velocity, is usually called the " blast area." With a larger area 
the pressure will be reduced, but with a smaller area will remain sub- 
stantially constant. The velocity of the discharge is practically that 
of the tips of the blades, whence the blast area is equal to 

' = 4.23 square feet, which, with this type of fan is found to be 

about J the projected rectangle of the wheel, therefore, 
The projected rectangle = 4.23 X 3 = 12.7 square feet. 
The proper width of periphery is found by dividing this area by the 

wheel diameter, thus, 

width of blades = ^- = 2.02 feet = 24.2 inches, 
6.3 

and speed of fan = — - =311 r.p.m. 

P 3.14X6.3 F 

w _ Volume of gas (cu. ft. per min.) X Pressure (lb. per sq. ft.) 
33,000 X efficiency of fan 

W = 26 > 0Q0 x 5 - 2 = io.2 brake horse power. 
33,000 X 0.4 

(5.2 = pressure in pounds per square foot equivalent to one inch of 
water, and 0.4 is the mechanical efficiency for 75 per cent opening as 
taken from curve in Fig. 137.) 

Assuming a steam consumption of 70 pounds per brake horse power 
for a small, simple non-condensing high-speed engine, the steam con- 
sumed per hour will be 

10.2 X 70 = 714 pounds per hour, or 2.3 per cent of the total steam 
capacity of the boilers. 

Table 34 gives the capacity and horse power required for various 
sizes of forced-draft fans, and Table 35 gives similar data for induced- 
draft fans. 



MECHANICAL DRAFT 



261 



154. Chimney vs. Mechanical Draft. — The choice of chimney or 
mechanical draft depends largely upon local conditions. Many power 
plants with tall stacks are provided with forced-draft apparatus to be 
used in emergencies, but as a general rule where ordinances require 
high chimneys mechanical draft is not considered. In a few isolated 
cases stokers of the forced-draft type are used in connection with 
chimneys as high as 250 feet, but such installations are rare, and not 
to be recommended. 

Where there are no limitations to the height of stack, mechanical 
draft offers many advantages over chimney draft, especially for rail- 
road work and large lighting plants. With certain types of grates and 
for low-grade fuels and anthracite culm or dust, it is indispensable. 
Again, where a fair quality of fuel is obtainable the size of plant may 
determine the choice. 

First Cost: In small plants of say 100 to 150 horse power the cost 
of a guyed steel chimney, 75 feet in height or less, would be consider- 
ably less than that of a mechanical-draft system, and once erected cost 
practically nothing for operation, while the power required to operate 
a fan in so small a plant would amount to 5 per cent or more of the 
total steaming capacity. 

TABLE 34. 

CAPACITIES OF FORCED-DRAFT FANS. 

{Power.) 













For Forced Draf 


,, Temperature of Air 60°. 










Cubic Feet 
of Air De- 


Pressure in Inches of Water. 


. 
















Diam- 
eter of 


livered to 
Furnace 


0.5 


0.75 


1.00 


1.25 


1.50 


2.00 


2.50 


































per 


S 


P4 


a 


Ph' 


§ 


Ph' 


a 


Ph° 


§ 


Ph' 


3 


Ph 


3 


Ph' 




Minute. 


Ph 

510 


1.6 


Ph 

Ph 

560 


1.8 


Ph 

600 


1.9 


Ph 
Ph~ 

640 


2.1 


Ph 
Ph" 

710 


w 

2.3 


Ph 
Ph 

780 


H 

2.5 


Ph 
PS* 

850 


w 


2' 6" 


4,200 


2.7 


3' 


5,800 


430 


2.2 


460 


2.4 


490 


2.6 


530 


2.8 


590 


3.1 


640 


3.4 


710 


3.8 


3' 6* 


7,800 


360 


3.0 


400 


3.3 


420 


3.5 


450 


3.8 


500 


4.2 


550 


4.6 


610 


5.1 


4' 


10,000 


320 


3.9 


350 


4.2 


370 


4.4 


400 


4.9 


440 


5.4 


480 


5.9 


530 


6.5 


4' 6" 


12,400 


290 


4.8 


310 


5.2 


330 


5.6 


360 


6.0 


400 


6.7 


430 


7.3 


470 


8.0 


5' 


15,200 


250 


5.9 


270 


6.4 


290 


6.8 


310 


7.4 


350 


8.2 


380 


8.9 


420 


9.8 


5' 6* 


18,200 


230 


7.0 


250 


7.7 


270 


8.2 


300 


8.8 


330 


9.8 


360 


10.6 


390 


11.8 


6' 


21,400 


210 


8.3 


230 


9.1 


250 


9.6 


260 


10.4 


290 


11.5 


320 


12.5 


350 


13.9 


7' 


28,800 


180 


11.2 


200 


12.2 


210 


13.0 


230 


14.0 


250 


15.5 


280 


16.8 


300 


18.7 


8' 


37,200 


160 


14.4 


170 


15.7 


190 


16.7 


200 


18.1 


220 


20.1 


240 


21.8 


270 


22.5 


9' 


46,800 


140 


18.1 


160 


19.8 


170 


21.1 


180 


22.7 


200 


25.3 


220 


27.4 


240 


30.3 


10' 


57,400 


130 


22.2 


140 


24.3 


150 


25.8 


160 


27.9 


180 


3.1 


200 


33.6 


210 


37.2 



Discharge velocity 2000 feet per minute. 



262 



STEAM POWER PLANT ENGINEERING 



TABLE 35. 
CAPACITIES OF INDUCED-DRAFT FANS. 

{Power.) 



For Induced Draft, Temp, of Flue Gases 500°. 





Cubic Feet 
of Air at 
60°Temp. 

Drawn 
into Fur- 
nace per 

Minute. 


Pressure in Inches of Water. 


Diam- 
eter of 


0.5 


0.75 


1.00 


1.25 


1.50 


2.00 


2.50 


Fan. 


3 

Oh' 

688 
580 
486 
432 
390 
337 
310 
283 
243 
216 
189 
175 


Oh' 

M 

2.2 

3.0 

4.0 

5.3 

6.5 

8.0 

9.5 

11.2 

15.1 

19.4 

24.4 

30.0 


3 

Oh' 

pi 

756 
621 
540 
472 
418 
364 
337 
310 
270 
230 
216 
190 


Ph* 

w 

2.4 

3.2 

4.5 

5.7 

7.0 

8.6 

10.4 

12.3 

16.5 

21.2 

26.7 

32.8 


Oh 
Ph 

810 
661 
567 
500 
445 
391 
364 
337 
283 
256 
230 
202 


Ph' 

w 

2:6 

3.5 

4.7 

6.1 

7.5 

9.2 

11.1 

13.0 

17.5 

22.5 

28.5 

34.8 


Ph' 
Ph- 

864 
715 
607 
540 
486 
418 
405 
351 
310 
270 
243 
216 


pj 
W 

2.8 

3.8 

5.1 

6.6 

8.1 

10.0 

11.9 

14.0 

18.9 

24.4 

30.6 

37.6 


3 

Ph 

pi 

958 
796 
675 
594 
540 
472 
445 
391 
337 
297 
270 
243 


Ph 
W 

3.1 

4.2 

5.7 

7.3 

9.0 

11.1 

13.2 

15.5 

20.9 

27.1 

34.1 

41.8 


d 

Ph 
Ph 

1053 

864 
742 
648 
580 
513 
486 
432 
378 
324 
297 
270 


Ph' 

a 

3.4 

4.6 

6.2 

8.0 

9.8 

12.0 

14.3 

16.9 

22.6 

29.4 

37.0 

45.3 


Ph' 

Ph 

1147 
958 
823 
715 
634 
567 
526 
472 
405 
364 
324 
283 


Ph" 
U 


2' 6* 
3' 

3' 6" 
4' 

4' 6" 
5' 

5' 6" 
6' 
V 
8' 
9' 
10' 


3,000 

4,200 

5,700 

7,300 

9,300 

11,100 

13,300 

15,600 

21,000 

27,100 

34,200 

41,900 


3.6 
5.1 

6.9 
8.8 
10.8 
13.2 
15.9 
18.7 
25.2 
30.4 
40.9 
50.2 



A tall, self-supporting chimney for larger plants, however, is very 
costly as compared with a fan system of equal capacity. For example, 
a brick chimney 175 feet high and 10 feet in diameter, foundation 
and all, capable of furnishing the necessary draft for a 3000-horse-power 
plant, will cost about $10,000. A two-fan induced system of equiv- 
alent capacity will cost in the neighborhood of $5000, a one-fan 
system $3500, and a forced-draft system $2500. See Fig. 140. With 
interest at 5 per cent, depreciation 5 per cent, taxes 1 per cent, and 
insurance one-half per cent, the annual fixed charges will be $575, 
$402.50, $287.50, respectively, for the fan equipment. 

Depreciation and Maintenance : The depreciation of a well-designed 
masonry or concrete stack is very low, and 2 per cent is a liberal factor. 
Maintenance is practically negligible, as it requires no attention what- 
ever for years. A steel stack, however, must be kept well painted or 
corrosion will take place rapidly. The depreciation and maintenance 
charges on a mechanical-draft system will range from 4 per cent to 10 
per cent of the original outlay. 

Cost of Operation: Once erected, the comparative cost of operating 
a chimney is practically nothing; that is, of course, on the assumption 
that the chimney and fan exhaust equal volumes of gas per pound of 



MECHANICAL DRAFT 



263 



fuel and at the same temperature. A fan system requires for its opera- 
tion from one and one-half per cent to five per cent of the total steaming 
capacity of the plant, depending upon the type and character of the 
fan engine or motor, and the conditions of operation. 

Efficiency: With fan draft a very thick fire can be maintained on 
the grate, thus permitting a high rate of combustion, and minimum 
air per pound of fuel, both of which result in increased boiler efficiency. 



15 


















































j 










































































14 






























































13 


























































































































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Horse Power 
Fig. 140. Comparative Costs of Chimneys and Mechanical Draft. (W. B. Snow.) 

The influence of the rate of combustion on air supply is illustrated in 
Fig. 141. For the same temperature of discharge each pound of air 
in excess of theoretical requirements results in a loss of about one per 
cent of the total heat in the fuel. (See Table 3.) With fan draft an 
average figure is 18 pounds of air per pound of bituminous coal against 
24 pounds for the chimney, a saving of 5 per cent in favor of the fan. 
Again, a fan permits of a low temperature of the flue gases without 



264 



STEAM POWER PLANT ENGINEERING 



affecting the draft, while lowering the temperature in the chimney 
reduces the draft as shown in Table 24. From Table 4 we see that 
a reduction in flue gas temperature of 25 degrees F. will increase the 
boiler efficiency about one per cent. With an economizer the flue 
gases may be reduced to 350 degrees F., with a net saving of about 
500 — 350 = 150, or 6 per cent of the total fuel. It is in this connection 
that the fan draft is peculiarly suitable. Of course, the chimney may 
be provided with an economizer, effecting the same reduction in tem- 

300 



=5 200 



100 



o 

o 



10 



20 



40 



50 



X,b.Coal Burned Per Sq.Ft.Grate Per Hr. 
Fig. 141. Influence of Rate of Combustion on Air Supply. — Forced Draft. 



perature, but its height must be made sufficiently great to overcome 
the additional resistance of the economizer and the reduction in tem- 
perature of the chimney gases. 

Flexibility : With a fan the draft may be readily regulated for 
sudden increased or decreased requirements, independent of the boiler 
performance. Damp and muggy days appreciably affect the draft of 
a chimney, as do adverse air currents and high winds. 

Smoke: Smokeless combustion is more readily effected with arti- 
ficial draft than with natural draft, as a thicker fire can be carried, and 
the correct proportion of air can be more readily adjusted. 

Comparative Tests of Chimney and Mechanical Draft: Power, July, 1901, p. 22; 
Eng. Rec, July 25, 1903, p. 102; Eng., U. S., May 1, 1899, p. 105; Engr., U. S., 
April 15, 1907. 

155. Balanced Draft. — Fig. 142 illustrates an application of the 
McLean " Balanced Draft " system to a water-tube boiler. The 
equipment consists of a blower, the speed of which is regulated 



MECHANICAL DRAFT 



265 




a 

CD 

0Q 



266 



STEAM POWER PLANT ENGINEERING 



by the steam pressure, so that the draft in the fire box is main- 
tained at approximately atmospheric pressure. The chief claims for 
this system are (1) the velocity of the gases over the tubes is reduced, 
and short circuiting is prevented; (2) the correct proportion of air to 
fuel is readily maintained; (3) infiltration of air through the setting 
is impossible, as the pressures are " balanced"; (4) sudden changes 
in load are correctly taken care of. Tests of the apparatus at the 
Fuller Building, New York, gave excellent results (Trans. A.S.M.E., 
26-641). 



CHAPTER IX. 

STEAM ENGINES. 

156. Introductory. — The reciprocating steam engine is the most 
widely distributed and generally adopted prime mover in the power 
world although its field of usefulness has been greatly encroached upon 
in recent years by the steam turbine and the gas engine. The steam 
turbine has practically superseded the piston engine for large steam 
electric plants, while in other fields the gas engine offers many advan- 
tages, but the reciprocating steam engine is still an important heat 
engine and will probably continue to be a factor in the power world 
for years to come. 

The type of engine best suited for a given installation is the one 
which delivers the required power at the lowest cost, measured in 
dollars and cents, taking into consideration interest on the investment, 
operating expenses, maintenance and depreciation. 

157. Ideal Engine. — The thermal efficiency of the steam engine 
is expressed by the ratio of the heat equivalent of the work done on 
the piston per unit of time, to the heat supplied. The degree of per- 
fection realized is ascertained by comparing the performance of the 
real engine with that of an ideally perfect engine, working between the 
same temperature limits. The theoretical limit of perfection is that 
defined by the Carnot cycle, the efficiency of which is represented 
by the equation 

E C = T '~ T \ (66) 

in which 

T x = the highest absolute temperature of the working fluid. 
T 2 = the lowest absolute temperature of the working fluid. 

The upper limit of temperature is that corresponding to boiler pressure, 
and the lower limit to that of the exhaust steam. Evidently the 
greater the temperature range the more nearly does the ideal efficiency 
approach unity, but with the present limits of temperature used in 
steam engines, it cannot exceed about 35 per cent. 

The nearest approach of any actual engine to the Carnot cycle is 
accomplished by the Nordburg system of progressive feed heating, 

267 



268 STEAM POWER PLANT ENGINEERING 

in which the feed water is successively heated from the receivers inter- 
mediate between each pair of cylinders. (Engineering News, May 4, 
1899, p. 283.) Table 36 gives the Carnot efficiencies of condensing and 
non-condensing engines for ordinary ranges of steam pressures. 

The Carnot cycle is theoretically impossible for an engine using 
superheated steam at constant pressure, and, in general, it is not very 
closely simulated by engines using saturated steam. It is, therefore, 
i more instructive to select an ideal cycle which 

\ more nearly represents the performance of the 

\ actual engine. The diagram representing the 

\ operation of this perfect engine is shown in 

\. Fig. 142a, and is called the non-conducting or 

^v Rankine cycle, ab represents the admission 

_^^ 6 - of dry steam from the boiler at pressure p x ', 

&; be is an adiabatic expansion to exhaust pressure 

p 2 ; cd represents the exhaust, and rfa is an adia- 

batic compression to the initial pressure. 

The heat necessary to raise the feed water from the temperature of 

exhaust, or ideal feed water temperature, to the temperature in the boiler 

and evaporate it into dry steam is 

#i - r x +. q ± - q 2 , (66) 

in which 

H t = quantity of heat supplied to the cylinder per pound of steam. 
r t = heat of vaporization at pressure p v 
q x = heat of the liquid at pressure p v 
q 2 = heat of the liquid at pressure p 2 . 

The heat, H 2 , exhausted from the cylinder and which must be with- 
drawn when it is condensed is 

H 2 = x 2 r 2 , (66a) 

in which 

x 2 = quality of the steam at pressure p 2 . 
r 2 = heat of vaporization at pressure p 2 . 

x 2 may be calculated by the aid of equation 

* 2 = | 2 (^ + 1 -0 2 )> . (66b) 



STEAM ENGINES 



269 




270 



STEAM POWER PLANT ENGINEERING 




STEAM ENGINES 271 

in which 

T 2 = absolute temperature of steam at pressure p 2 . 
T x = absolute temperature of steam at pressure p v 
6 X = entropy of the liquid at pressure p v 
d 2 = entropy of the liquid at pressure p 2 . 
Other notations as above. 

The heat changed into work per pound of steam is 

H i ~ H 2 = r x + q x - q 2 - x 2 r 2 , (67) 

and the efficiency, E r , of the cycle is 

E = H >- H > = 'i + gi ~ v, - g, . ' (67a) 

H t r 1 + q 1 - q 2 

The steam consumption W, or water rate, lbs. per h.p. hr. of the perfect 
engine, may be expressed 

W= , 2545 (67b) 

n + 2i ~ g 2 . 

If the steam entering the cylinder is wet and of quality x lf substitute 
x 1 r 1 in above equations for r v 

If the steam is superheated at admission but becomes moist at the 
lower pressures, which is the usual case, the efficiency may be expressed 

E r i + gi +c 1 t s -x 2 r 2 -q 2 t 
r i + Qi + Cits - q 2 
in which 

c x = mean specific heat of the superheated steam at pressure p v 
t s = degree of superheat or difference in temperature between the 
superheated and saturated steam at pressure p v 

x 2 may be calculated by the aid of equation 



Jr 



Tc ~P + ^+e l = ^ + d 2 , (67d) 

in which 

t and T = thermometric and absolute temperatures of the superheated 
steam. 
c = true specific heat of superheated steam at temperature t. 

Other notations as above. 



272 



STEAM POWER PLANT ENGINEERING 




O 
d 

a 

o 

O 



I 

d 
d 



STEAM ENGINES 273 

For many purposes equation (67d) may be expressed 

e, log.|j + £+*,- 3& +0,. (67e) 

For highly superheated steam in which the steam is still superheated 
at exhaust 

T? — r i "^" ?1 ~^" C J S ~ r 2 ~ ?2 ~ C 2^S C67f ^ 

r t + q t + c t t s - q 2 
in which 

c 2 = mean specific heat of the superheated steam at exhaust. 

t/ = degree of superheat at exhaust. 

t s ' may be calculated by the aid of equation 

f T cdt ^r, n CW c'dt , r 2 ' 

Problems connected with the Rankine cycle may be conveniently 
solved by temperature-entropy tables to be found in connection with 
the usual steam tables or by the Mollier diagram as described in 
Appendix H. 

158. Thermal Efficiency of the Actual Engine. — In calculating the 
thermal efficiency of the real engine the heat supplied is reckoned above 
the sensible heat of the exhaust, thus, 

_ Heat converted into useful work 
Heat supplied 



42.42 



B.T.U. supplied per I.H.P. per minute 
42.42 



w(x l r l +q v - q 2 )' 
in which 



(69) 
(70) 



w = the weight of steam supplied to the engine per indicated horse 
power per minute, or per brake horse power per minute, de- 
pending upon whether the efficiency is to be referred to the 
indicated or to the brake horse power of the engine. 

Other notations as in (66) and (67). 

The figure obtained by dividing the efficiency of the real engine by 
that of the ideal engine is called the efficiency ratio, and is a measure of 
the extent to which the theoretical possibilities are realized. 



274 



STEAM POWER PLANT ENGINEERING 



The efficiency ratio is calculated on the basis of the indicated horse 
power or the developed horse power: 



Eff. Ratio = 



42.42 



E r w (H 1 - H 2 ) 



(71) 



This ratio expressed in different terms has been referred to as a Poten- 
tial Efficiency" by C. V. Kerr (Trans. A.S.M.E., 25-920), and as 
" Cylinder Efficiency " by Professor Reeve. 

The commercial economy of an engine is measured by the cost of 
producing power, and does not necessarily depend upon its thermal 
efficiency. The performances of steam engines are frequently stated 
in terms of (a) pounds of steam utilized per horse-power hour, (6) 
pounds of coal per horse-power hour, (c) cost in cents per horse-power 
hour, (d) B.T.U. per horse-power hour. From a commercial stand- 



TABLE 36. 
STEAM-ENGINE EFFICIENCIES. 

(Saturated Steam.) 





Non-Condensing; Back Pressure 


14.7 


Condensing; Back Pressure 1 Pound 






Absolute. 






Absolute. 




Gauge 


















Fress. 








Ratio 








Ratio 




Carnot 


Rankine 


Actual * 


b 
a 


Carnot 


Rankine 


Actual 


b 
a 




Cycle. 


Cycle (a). 


(b). 


Cycle. 


Cycle (a). 


(b). 










% 








% 


25 


7.5 


7.3 


5.5 


76.0 


22.6 


21.0 


11.6 


55.0 


50 


11.2 


10.7 


8.5 


80 


25.7 


23.5 


13.5 


60 


75 


13.7 


13.0 


10.4 


80 


27.8 


25.3 


15.9 


61 


100 


15.7 


14.8 


12.0 


81 


29.5 


26.7 


20.2 


76 


125 


17.3 


16.3 


13.5 


83 


30.8 


27.8 


20.3 


74 


150 


18.7 


17.5 


14.3 


82 


32.0 


28.8 


21.6 


75 


175 


19.8 


18.5 


14.8 


80 


32.9 


29.6 


21.9 


74 


200 


20.8 


19.3 


15.2 


79 


33.7 


30.2 


22.6 


75 


225 


21.6 


19.9 


15.5 


78 


34.5 


30.6 


22.6 


74 


250 


22.4 
23.0 
23.6 


20.5 

21.0 

. 21.4 






35.1 
35.6 
36.0 


31.0 
31.3 
31.5 






275 










300 





















* Best recorded performance of the actual engine, 1907 



point the cost of producing power is the most important basis of com- 
parison, but the latter expression is most satisfactory for scientific 
purposes, since it gives a basis of comparing the performances of all 
types of heat engines. 



STEAM ENGINES 



275 



159. Mechanical Efficiency. — The power of an engine may be 
expressed in terms of indicated horse power, brake horse power, 
or pump horse power, according to the class of engine. The ratio 
of the brake to the indicated power is the mechanical efficiency 
of the engine, the ratio of the electric horse power to the indicated 
power is the mechanical efficiency of the engine and generator com- 
bined, and the ratio of the pump horse power to the indicated power 
of the steam engine is the mechanical efficiency of the engine and pump 



45 



























n 




r 










X ^-< 






.— — ° 










C 










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*f 


r<& * 




•*o 


o 




















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% 


V 


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: 


























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/ 












Mechanical 

Efficiencies of 

75 KW.Generating Set 

Engine, Simple High Speed 

!Non Condensing 







// 


( 














/ / 




/o 














/ 




























































10 20 30 40 50 60 70 80 90 100 110 120 
Per Cent of Rated Load 

Fig. 143. 



140 150 



combined. The percentage of work lost in friction is therefore the 
difference between 100 per cent and the mechanical efficiency. 

Table 37 shows the mechanical efficiencies for several types of en- 
gines, and Fig. 143 the combined efficiency of a direct-connected high- 
speed engine and generator. (See Engine Friction, par. 167.) 

Mechanical Efficiency: Peabody, Thermodynamics, p. 430; Spangler, Steam 
Engineering, p. 205; Ripper, Steam Engine, p. 275; Ewing, Steam Engine, p. 186. 



The following numerical example will illustrate the calculation of 
the various efficiencies mentioned: 

A simple high-speed engine uses 30 pounds of steam per I.H.P. 
hour; initial pressure 100 pounds per square inch, gauge; exhaust 
pressure, atmospheric; I.H.P., 120; D.H.P., 102. Steam assumed to 
be dry and saturated at throttle. Required: (1) the actual thermal 
efficiency; (2) the efficiency of the Carnot cycle; (3) the efficiency of 
the Rankine cycle; (4) the efficiency ratio; (5) the mechanical 
efficiency. 



276 



STEAM POWER PLANT ENGINEERING 



TABLE 37. 

MECHANICAL EFFICIENCIES OF ENGINES. 



Kind of Engine. 


Horse Power. 


Efficiency at Full 
Load. 


Simple : 

1. High-speed, non-condensing 


150 
170 
275 


95 5 


2 High-speed, condensing 


96 


3. Low-speed, non-condensing 


94 


4. Low-speed, condensing 




Compound : 

5 High-speed, non-condensing 


150 

160 

900 

1000 

865 
712 


94 


6. High-speed, condensing 


98 


7. Low-speed, non-condensing 

8. Low-speed, condensing 


95 
95 


Triple: (combined efficiency of engine and 
pump) 
9 Pumping engine 


97.4 


Quadruple: (combined efficiency of engine and 
pump) 
10. Pumping engine 


93 







1. Buffalo Simple engine, 12 X 12, Elec. World, Sept., 1904, p. 147. 

2. Reeves Simple engine, 15 X 14, Elec. World, Oct. 1, 1904, p. 587. 

3. 24 X 48 Hamilton Corliss at Armour Inst, of Tech., 1898. 
4. 

5. Reeves Compound; Eng. Rec, July 1, 1905, p. 24. 

6. Reeves Compound; Eng. Rec. 

7. 21, 41 X 30 Cross Compound Ball & Wood, West Albany Station, N.Y.C. & H.R.R, 

8. 20, 40 X 42 Rice & Sargent; A.S.M.E., 29-1276. 

9. Allis Pumping Engine; Power, May, 1906, p. 299. 

10. Nordburg Pumping Engine; Eng. News, May 4, 1899, p. 280. 



(1) The actual thermal efficiency is 

42.42 



E = 



W {x 1 r 1 + q l - q 2 ) 
42.42 



30/60 (879.6 + 308.9 - 180.3) 
= 0.084. 

That is, only 8.45 per cent of the heat supplied to the cylinder above 
the temperature of the exhaust steam is converted into work. The 
assumption that the exhaust steam may be used to heat the feed water 
to its own temperature justifies reckoning the heat supplied above the 
exhaust temperature. 

(2) The efficiency of the Carnot cycle is 

T t -T 2 



E c 



798.8 - 



673 



798.8 



= 0.157, 



STEAM ENGINES 



277 



which means that if the engine were a perfect one employing the Car- 
not cycle between the same extremes of temperature as the actual 
engine, 15.7 per cent of the energy supplied would be converted into 
useful work. 

(3) On the basis of the Rankine cycle the ideal efficiency becomes 



V. — X i r i "*" 9l ^2 r 2 9.2 

x Si + 9i ~ 92 



879.6 + 308.9 - 0.885 X 969.7 - 180.3 



879.6 + 308.9 - 180.3 



= 0.149, 



the numerator representing the heat converted into work per pound of 
steam and the denominator the heat supplied. Thus if the engine be 



300 
275 

250 

<v 
60 2-25 

"5 
O 
a r 200 

d 1 

co 175 
o 
Ph 
.£150 

Hi 

oT 
£125 

03 

in 
P 

£ 100 























































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I 


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8 




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i 


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7 1 


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7 
























ft 


































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v 


f 


V 


J 








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c 

g 


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/ 


// 




V 








w 
























A 








4 

J*/ 
















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Status of the 
Piston Engine 

1907 
Saturated Steam 




? 








V 










%y 







































8 10 12 14 16 18 20 22 24 26 
Thermal Efficiency, Per Cent 

Fig. 144. 



32 34 



assumed to have a non-conducting cylinder and work on the Rankine 
cycle, it would be capable of utilizing 14.9 per cent of the heat sup- 
plied. 

(4) The efficiency ratio is expressed: 



Efficiency ratio = — ~ = ^— - 
J E r 14.9 



56.7 per cent, 



278 STEAM POWER PLANT ENGINEERING 

which indicates the degree of perfection of the engine or the extent to 
which it realizes the maximum efficiency theoretically possible. The 
weight of steam which would have to be consumed per horse power 
per hour to actually attain the Rankine efficiency would be 

30 X ^2^ = 30 x 0.567 = 17 pounds. 

T) TT "P 1 f)9 

(5) The mechanical efficiency = ' ' ' = =0.85 = 85 per cent. 

The relation between the actual and the theoretical efficiencies 
based upon the Rankine cycle, and upon the best recorded perform- 
ances of modern engines, using saturated steam, is shown graphically in 
Fig. 144. (Also see Tables 39 and 40.) The theoretical curves are 
calculated upon the assumption of complete expansion, the back 
pressure being 14.7 pounds gauge for non-condensing and one pound 
absolute for condensing engines. 

The highest recorded (1907) efficiency ratio for a steam engine 
using saturated steam is 83 per cent non-condensing and 76 per cent 
condensing (see Table 39), and 78.5 per cent for a condensing engine 
using superheated steam (Table 43). 

160. Heat Losses in the Steam Engine. — The principal losses which 
tend to lower the efficiency of the steam engine are due to 

(a) Presence of moisture in the steam at admission. 

(b) Leakage past valves and piston. 

(c) Cylinder condensation. 

(d) Clearance volume. 

(e) Incomplete expansion. 
(/) Wire drawing. 

(g) Friction. 
(h) Radiation. 

161. Moisture. — The presence of moisture in the steam pipe is 
due to condensation caused by radiation or to priming at the boiler. 
Unless removed by some separating device between boiler and engine 
the amount of moisture entering the cylinder may be from 1 to 5 per 
cent of the total weight of steam, and the work done per pound of 
fluid is correspondingly reduced. This loss should not be charged 
against the engine, however, and its performance should be reckoned 
on the dry steam basis. Experiments reported by Professor R. C. Car- 
penter (Trans. A.S.M.E., 15 — 438) in which water in varying quantities 
was introduced into the steam pipe, causing the quality of the steam 
to range from 99 per cent to 57 per cent, showed that the consumption 



STEAM ENGINES 279 

of dry steam per I.H.P. hour was practically constant, the water acting 
as an inert quantity. An efficient separator will remove practically 
all the entrained water. 

Influence of Moisture on Steam Economy: Trans. A.S.M.E., 18-699; Benjamin, 
Heat and Steam; Perry, Steam Engine, p. 353; Rankine, Steam Engine, p. 407; 
Ripper, Steam Engine, p. 33; Ewing, The Steam Engine, p. 139. 

162. Leakage of Steam. — The loss due to leakage is a variable 
factor depending upon the design and condition of the engine, and is 
greater with saturated than with superheated steam. The usual method 
of measuring leakage past the valves and piston while the engine is at 
rest is likely to give erroneous results as demonstrated by Callender 
and Nicolson (Peabody, " Thermodynamics/' p. 351) in tests made on a 
high-speed automatic balanced valve engine and on a quadruple expan- 
sion engine with plain unbalanced slide valves. With the engines at 
rest they found that the leakage past valves and piston was insignificant, 
but when in operation the leakage from the steam chest into the exhaust 
was very considerable indeed. It was thought that a large proportion 
of the leakage was probably in the form of water formed by condensa- 
tion of steam on the seat uncovered by the valve. 

According to the report of the Steam Engine Research Committee 
(Eng. Lond., March 24, 1905, p. 298), leakage through a plain slide 
valve is independent of the speed of the sliding surfaces, and directly 
proportional to the difference in pressure on the two sides; with well- 
fitted valves the leakage is never less than 4 per cent of the volume of 
steam entering the cylinders, and is often greater than 20 per cent. 

Peabody, Thermodynamics, p. 350; Eng. Rec., May 22, 1897, p. 529; Barms, 
Engine Tests, p. 251. 

163. Cylinder Condensation. — A large percentage of the steam 
admitted to the cylinder is condensed, due to the absorption of heat 
by the relatively cool cylinder walls. Condensation continues during 
expansion until the temperature of the steam falls below that of the 
metal, when the process is reversed and a part of the moisture is re- 
evaporated. Unless the cylinder is one of a compound series, the heat 
absorbed from the cylinder walls during exhaust does no useful work 
and is lost. The condensation up to the point of cut-off (initial con- 
densation) may amount to from 15 to 30 per cent, and is often as high 
as 50 per cent of the total weight admitted to the cylinder. The 
initial condensation becomes greater as the difference between initial 
and exhaust pressures is increased, and diminishes as the speed of the 
engine increases. Cylinder condensation and leakage are ordinarily 



280 



STEAM POWER PLANT ENGINEERING 



classified together, as there is no way of separating them accurately. 
They represent that part of the feed water which is not accounted for 
by the indicator diagram. 

Tests of 20 simple high-speed engines by G. H. Barms, Fig. 145, 
show some results obtained for various percentages of cut-off. Also 
see table compiled by C. H. Peabody, " Thermodynamics of the Steam 



60 




o 






























































1 
















Condensation.and Leakage 
for 
Simple Engines 

using 
Saturated Steam* 
















o 














1 

DO 










\o 



































s 






















a 
o 










o 





n 


s° 
































3 s 


^ 


< 














J? 

1 20 






















>r 


^ 


3 r 






































O 


io 






















































E 


lgine 


Tests 


, Ban 


us, p. 


254. 



10 15 20 35 

Percentage of Cut Off 

Fig. 145. 



40 



Engine," p. 336, showing analysis of the heat interchanges for a number 
of different types of steam engine. 

The various heat losses, including cylinder condensation and leakage, 
are best determined by transferring the indicator diagram to the tem- 
perature entropy or 0<f> chart. (See Appendix C.) This is useful for 
certain scientific investigations, but is unnecessary for commercial tests. 

Cylinder Condensation : Trans. A.S.M.E., I, 184, III, 215, IV, 88, VII, 375, 
XVIII, 950; Cotterill, Steam Engine, p. 331; Spangler, Steam Engineering, p. 228; 
Thurston, Manual of the Steam Engine, I, 271, 488, 585; Heck, Steam Engine, 
pp. 109, 113, 119; Ewing, Steam Engine, p. 148; Hutton, Heat Engines, p. 319; 
Peabody, Thermodynamics, pp. 241, 412; Reeve, Thermodynamics, p. 198; Wood, 
Thermodynamics, p. 212; Perry, Steam Engine, p. 78; Ripper, Steam Engine, p. 25. 

Initial Condensation: Cotterill, Steam Engine, p. 274; Golding, 6$ Diagram, 
p. 63; Marks, Steam Engine, p. 195; Peabody, Thermodynamics, p. 359; Popplewell, 
Heat Engine, pp. 323, 351; Reeve, Thermodynamics, pp. 156, 221; Ripper, Steam 
Engine, pp. Ill, 168. 



STEAM ENGINES 281 

Condensation during Expansion: Trans. A.S.M.E., III, 286; Hutton, Heat 
Engines, pp. 223, 286; Pupin, Thermodynamics, p. 88; Rankine, Steam Engine, 
p. 385; Reeves, Thermodynamics, p. 221. 

Entropy : Power, Jan. 21, 1908; Baynes, Thermodynamics, p. 94; Benjamin, 
Heat and Steam, p. 37; Berry, Temperature-Entropy Diagram ; Boulvin, Entropy 
Diagram; Ewing, Steam Engine, p. 103; Golding, 0<f> Diagram; Hutton, Heat 
Engines, p. 276; Peabody, Thermodynamics, p. 97 ; Reeve, Thermodynamics, p. 39; 
Swinburne, Entropy; Wood, Thermodynamics, p. 136: Heck, Steam Engine, 
Chap. VI. 

164. Clearance Volume. — The portion of the cylinder volume not 
swept through by the piston but which is nevertheless filled with 
steam when admission occurs is called the clearance volume. It is the 
space between the end of the piston when on dead center and the 
inside of the valves covering the ports. It varies from about 1 per cent 
of the piston displacement in very large engines with short steam 
passages to 10 per cent or more in small high-speed engines. When 
the steam retained in the clearance space is compressed to the initial 
pressure and expansion is carried down to the back pressure, the clear- 
ance has little effect upon the economy of the engine, but since expan- 
sion and compression are seldom complete in actual practice, the loss 
may be considerable. (Ripper, " Steam Engine," p. 103.) The shorter 
the cut-off the greater will be the ratio of the weight of cushion steam 
to that of the steam supplied and hence the greater the relative loss. 
In large slow-speed engines the loss may be insignificant if the clearance 
volumes are small, while in small high-speed engines it may be con- 
siderable. 

The ratio of expansion is decreased by clearance; for example, an 
engine cutting off at one-fifth, neglecting clearance has an apparent 
ratio of expansion of 5, but if the clearance volume is 10 per cent the 
actual ratio is only 3.66. One of the few recorded tests relative to the 
influence of clearance on the economy of a high-speed engine was con- 
ducted on a 14x15 Allfree engine. (Power, May, 1901.) With a 
clearance volume of 2.2 per cent, initial pressure 105 pounds gauge, 
and 172 r.p.m., the best performance was 23.7 pounds of dry steam 
per I.H.P. hour. With the same steam pressure and speed, but 
with clearance volume increased to 6 per cent by the use of 
a shorter piston, the best performance was 28.3 pounds per I.H.P. 
hour. In both cases the compression was carried up to admission 
pressure. 

Loss by Clearance in Steam Engines ; Trans. A.S.M.E., XVIII, 176. Clearance in 
Compound Engines: ibid., I, 173. Clearance in M ulti-cy Under Engines : ibid., XI, 151. 
Effect of Clearance: Ewing, Steam Engine, p. 145; Hutton, Heat Engines, p. 334; 



282 



STEAM POWER PLANT ENGINEERING 



Peabody, Thermodynamics, p. 407; Popplewell, Heat Engines, p. 332; Reeve, Ther- 
modynamics, pp. 223, 256; Ripper, Steam Engine, pp. 63, 101; Spangler, Steam 
Engineering, p. 114; Wood, Thermodynamics, p. 197. 

165. Loss Due to Incomplete Expansion and Compression. — Theo- 
retically the loss due to incomplete expansion is considerable. For 
example, the theoretical steam consumption of a perfect engine (Ran- 
kine cycle) expanding from 120 pounds absolute to a condenser pres- 
sure of 2 pounds absolute is 9.6 pounds per horse-power hour. If the 
expansion were carried to only 5 pounds absolute, the exhaust pressure 
remaining the same, the steam consumption would be increased to 
11.8 pounds per horse-power hour, a difference of 22 per cent for an 
increase in terminal pressure of only 3 pounds per square inch. The 
theoretical water rates for various terminal pressures are given below. 



Terminal Pressure, 

Pounds per Square 

Inch Absolute. 


Steam Consumption 
of Perfect Engine. 


Terminal Pressure, 

Pounds per Square 

Inch Absolute. 


Steam Consumption 
of Perfect Engine. 


1 
1.5 

2 
2.5 


8.5 
9.1 
9.6 
10 


3 
4 
5 
6 


10.4 
11.1 
11.8 
12.3 



In actual engines expansion is seldom complete, since it would 
necessitate increased bulk and weight of engine, and the work done 
by the steam in the last stages would not compensate for the increased 
cost. 

In single-cylinder engines maximum economy is effected when the 
terminal pressure is considerably above that of the exhaust, since the 
gain due to complete expansion is more than offset by the increased 
cylinder condensation. This is true to a certain extent in all engines 
irrespective of the number of cylinders. Tests by G.H. Barrus ("Engine 
Tests," 1900) to determine the terminal pressures effecting maximum 
economy for various types of engine gave results as follows : 



Simple slide-valve engines, non-condensing 
Simple slide-valve engines, condensing .... 
Simple Corliss engines, non-condensing. . . . 

Simple Corliss engines, condensing 

Compound engines, non-condensing 

Compound engines, condensing 



Terminal Pressure, 


Pounds Absolute. 


30 to 40 


25 to 30 


20 to 25 


15 to 18 


18 to 22 


3 to 5 



STEAM ENGINES 



283 



In high-speed engines a certain amount of compression is desirable 
for its cushioning effect; outside of this mechanical feature compression 
may or may not be of benefit to the engine as will be explained later. 
Zuener in his treatise on theoretical thermodynamics proves deduc- 
tively that in an engine with a large clearance volume the loss due to 
clearance is completely eliminated if the compression is carried up to 
admission pressure, a conclusion which tests by Jacobus, Carpenter, 
and others fail to confirm. A series of tests by Professor Jacobus 
(Trans. A.S.M.E., 15-918) on a 10x11 high-speed automatic engine 
at Stevens Institute show decreasing economy with increase of com- 
pression, the initial pressure, cut-off, and release remaining constant. 
The results were as follows : 



Proportion of initial pressure up 

steam is compressed 

Steam, pounds per I.H.P. hour. . . . 



to which the 




Full 
38 



Tests by Carpenter (Trans. A.S.M.E., 16-957) on the high-pressure 
cylinders of the Corliss engine at Sibley College gave: 



Compression, per cent 

Brake horse power 

Steam, pounds perB.H.P. hour. 



11.4 


25 


30 


29 


33 


33.3 



35.2 

26 

34 



6.8 

6.6 
6.5 
6.4 
6.3 
6.2 
6.1 
6.0 
5.9 
5.8' 
5.7 
5.6 
5.5 














. 








48 
47 
46 
15 














100 Lb. Gauge 


^^ 


\ 






















;^w 












^^"t 
















< 










H 












^* 


^vV 








13 






















n 










%^ 












li 






^ 




^*Oj 


e ^ 










•10 














\^ 








39 

38 
37 
,36 
35 


s* 




Influence of Back 

Pressure on the Economy of 

an 8 x 10 Automatic High Speed 

Non Condensing Engine 


































1 

1 1 








*> 



6 8 10 12 14 

Back Pressure.Q). Per Sq.In.Gauge 

Fig. 145a. 



16 



18 



Opposed to these figures are tests which show an improvement in 
economy when compression is increased. 

Fig. 145a shows the influence of increasing back pressure on the 



284 STEAM POWER PLANT ENGINEERING 

economy of an 8 x 10 automatic high-speed engine at Armour Institute 
of Technology. 

Cut-off. — Best for Different Pressures; Trans. A.S.M.E., 4-89; In Compound 
Engines: ibid., p. 549; Most Economical Point of: ibid., 8-486; Hutton, Heat 
Engines, p. 232; Peabody, Thermodynamics, p. 210; Spangler, Steam Engineering, 
p. 109; Wood, Thermodynamics, pp. 200, 433; Klein, High-Speed Engines, p. 7; 
Ewing, Steam Engine, p. 84. 

Ratio of Expansion: Trans. A.S.M.E., 2-19, 128, 10-576, 11-166; Ewing, Steam 
Engine, pp. 47, 159; Wood, Thermodynamics, pp. 154, 172, 197, 295; Rankine, Steam 
Engine, pp. 378, 553; Reeve, Thermodynamics, p. 228; Thurston, Manual of the 
Steam Engine, 1-271, 725, 2-14. 

Incomplete Expansion: Boulvin, Entropy Diagram, p. 28; Cotterill, Steam Engine, 
p. 240; Perry, Steam Engine, p. 364; Popplewell, Heat Engines, p. 332; Peabody, 
Thermodynamics, p. 238; Reeve, Thermodynamics, p. 222; Heck, Steam Engine, 
p. 78. 

Compression. — Efficiency of Compression in Steam Engine : Engr., Lond., Nov. 3, 
1905, p. 434; Compression as a Factor in Steam Engine Economy : Trans. A.S.M.E., 
14-1067; Effect of Compression on Water Consumption: ibid., 15-815; Engine Com- 
pression: ibid., 7-708; In High-Speed Engines: ibid., 7-202; In Steam Cylinder: ibid., 
2-341. 

Back Pressure. — Back Pressure as Modifying Economy : Trans. A.S.M.E., 18-283; 
On Valves : ibid., 3-150; General : Ewing, Steam Engine, pp. 84, 145; Klein, High- 
Speed Engines, p. 11; Perry, Steam Engine, p. 75; Reeve, Thermodynamics, p. 223; 
Ripper, Steam Engine, p. 53. 

166. Loss due to Wire Drawing. — Wire drawing, or the drop in 
pressure due to the resistances of the ports and passages, has the effect 
of reducing the output of the engine to some extent, since the pressure 
within the cylinder is less than that at the throttle during admission 
and greater than discharge pressure at exhaust. The steam may be 
dried to a small extent during admission. In single- valve engines the 
effects of wire drawing are decidedly marked and the true points of 
cut-off and release are sometimes difficult to locate on the indicator 
card. In engines of the Corliss or gridiron-valve type the effects are 
hardly noticeable. 

Wire Drawing: Trans. A.S.M.E., 2-344, 1-174; Ewing, Steam Engine, pp. 95, 
143, 207; Boulvin, Entropy Diagram, p. 56; Popplewell, Heat Engine, p. 320; 
Rankine, Steam Engine, p. 413; Reeve, Thermodynamics, pp. 105, 221; Ripper, 
Steam Engine, p. 73; Wood, Thermodynamics, p. 195; Heck, Steam Engine, pp. 183, 
224, 230. 

167. Loss due to Friction of the Mechanism. — The difference between 
the indicated horse power and that actually developed is the power 
consumed in overcoming friction, and varies from 4 to 20 per cent of 
the indicated power, depending upon the type and condition of the 
engine. Engine friction may be divided into (1) initial or no-load 



STEAM ENGINES 



285 



friction and (2) load friction. The stuffing-box and piston-ring friction 
is practically independent of the load, while that of the guides, bearings, 




25 



50 



100 125 150 175 200 225 250 2T5 
Developed Horse-Power 



Fig. 146. Typical Curves of Steam Engine Friction. 

and the like increases with the load. In Fig. 146, curve A gives the 
relation between the frictions for a four-slide-valve horizontal cross 
compound engine, and B that for a simple non-condensing Corliss. 



TABLE 38. 
DISTRIBUTION OF FRICTION IN SOME DIRECT-ACTING STEAM ENGINES. 

(Thurston.)* 





Percentage of Total Engine Friction. 


Parts of Engines where Friction 
is Measured. 


" Straight 
Line " 
Balanced 
Valve. 


" Straight 
Line " 
Unbalanced 
Valve. 


Traction 
Engine 
Locomotive 
Valve Gear. 


Automatic 

Balanced 

Valve. 


Condensing 
Engine 
Balanced 
Valve. 


Main bearings 


47.0 


35.4 


35.0 


41.6 


46.0 


Piston and piston rod 


32.9 


25.0 


21.0 


49.1 




Crank pin 


6.8 


5.1 


13.0 


21.8 






Crosshead and wrist pin. ...... 


5.4 


4.1 




Valve and valve rod 


2.5 


26.4 


22.0 


9.3 






21.0 


Eccentric strap 


5.4 


4.0 






Link and eccentric 






9.0 
















Air pump 










12.0 
















100.0 


100.0 


100.0 


100.0 


100.0 



* " Friction and Lost Work in Machinery," p. 13. 



286 



STEAM POWER PLANT ENGINEERING 



(Peabody's " Thermodynamics," pp. 433 and 437.) Curve C is plotted 
from the tests of a Reeves vertical cross, compound condensing engine 
(Engineering Record, July 1, 1905, p. 24), and D from the test of an Ames 
simple high-speed non-condensing engine. (Engineering Record, Vol. 27, 
p. 225.) A large number of recorded tests show less friction at full load 
than at no load, but this is probably due to error or to variations 
in lubrication. With first-class lubrication it is usually sufficiently 
accurate to assume the friction to be constant and equal to the initial 
friction at zero load. The distribution of the frictional losses in a 
number of engines is given in Table 38. 

Friction in Engines : Trans. A.S.M.E., 8-86, 10-10, 8-108, 9-74, 82, 7-639, 641, 
1-153; Ewing, Steam Engine, p. 186; Ripper, Steam Engine, p. 275; Peabody, 
Thermodynamics, p. 430; Perry, Steam Engine, p. 270; Ripper, Steam Engine, 1-540; 
Heck, Steam Engine, 316-318. 

The efficiency of the fluid in the steam engine cylinder may be in- 
creased by (1) raising the boiler pressure, (2) compounding, (3) use of 
reheater-receiver, (4) steam jacketing the cylinders, (5) increasing the 
rotative speed, (6) superheating, (7) diminishing the back pressure by 
securing a more perfect vacuum. 

168. Effect of Increased Steam Pressure. — A consideration of the 
Rankine and Carnot cycles indicates that theoretically the greater the 
temperature range the greater will be the efficiency. (See Table 36.) 
In the actual engine the temperature range is most readily increased 
by raising the boiler pressure, since the limit of the back pressure is 
practically fixed by the cooling medium in the condenser. The theoreti- 
cal gain resulting from increased pressure range is, however, very con- 
siderably affected by the increased losses due to cylinder condensation. 

Fig. 147 shows the results of tests made at the Armour Institute of 
Technology on an 8 x 10 automatic high-speed piston-valve engine, 
showing marked gain with increase of initial pressure up to a certain 
point when the condensation losses became sufficiently great to neu- 
tralize the advantage which would otherwise be gained. 

The following figures were obtained in tests of a small Willans engine, 
non-condensing, under different steam pressures : 



Initial Pressure, Gauge. 


Pounds Steam per I.H.P. 
Hour. 


B.T.U. per I.H.P. per 
Minute. 


36.3 


42.8 


700 


51.0 


36.0 


595 


74.0 


32.6 


544 


85.0 


29.7 


495 


97.0 


26.9 


450 


110.0 


27.8 


465 


122.0 


26.0 


436 



STEAM ENGINES 



287 



Referring to Table 36, it may be noted that both the theoretical 
and the actual efficiencies increase very slowly for pressures above 150 
pounds. Practically, gain in efficiency due to increasing the pressure 



6.2 



6.1 



S 6.0 

f 
| 6.0 

B 

3 5.8 

a 



5.7 



5.6 



5.5 



















- 




V x 












•^ 








>,* 


















^ 


N^K 


y/ 
















4^ 


\ 














-* 


ps 


% 


V 






- 


















_ 


• 














* 


"^ 



48 



47 W 

w 



44 



43 0, 



75 



Fig. 147. 



85 



90 95 100 105 

Initial Gauge Pressure.Lb.per Sq.In. 



110 



115 



120 






Influence of Initial Pressure on the Economy of a Small, High-Speed, 
Non-Condensing Engine. 



above about 200 pounds is at the expense of increased first cost and 
maintenance and is only resorted to when small weight and space are 
the most important considerations. 

The range of pressures sanctioned by modern practice for different 
types of engines is as follows: 



Type of Engine. 



Simple slow speed 

Simple high speed 

Compound high speed, non-condensing 

Compound high speed, condensing 

Compound slow speed, condensing 

Triple expansion, condensing 

Quadruple expansion, condensing 



Range in Pres- 
sure (Gauge). 



60-120 
70-125 
100-170 
100-160 
125-200 
140-210 
125-225 



Average. 



90 
100 
130 
125 
150 
175 
200 



Steam Pressure : Trans. A.S.M.E., 4-88, 5-269, 6-572; Peabody, Thermo- 
dynamics, p. 248; Ripper, Steam Engine, p. 306; Engine Tests, Barrus, p. 258. 

169. Receiver-Reheaters. — The receivers between the cylinders of 
multi-expansion engines are frequently equipped with heating coils 
as illustrated in Fig. 332, the function of which is to superheat the 
exhaust steam before delivering it to the cylinder immediately follow- 



288 STEAM POWER PLANT ENGINEERING 

ing, with a view of reducing the losses occasioned by cylinder conden- 
sation. The coils are supplied with live steam under boiler pressure 
and may serve to evaporate a portion of the moisture or to actually 
superheat the steam supplied to the following cylinder. The question 
of the propriety of using reheaters is an open one, since reliable data 
relative to their use are meager and discordant. The conditions under 
which the few recorded tests were made are too diverse to warrant 
definite conclusions. Some show an appreciable gain in economy, 
others a decided loss. A reheater is of little value in improving the 
thermodynamic action of the engine, and is probably a loss unless it 
produces a superheat of at least 30 degrees F., and to be fully effective 
should superheat above 100 degrees F, (L. S. Marks, Trans. A.S.M.E., 
25-500.) The effectiveness of the reheater will evidently be increased 
by the removal of the greater portion of the moisture from the exhaust 
steam before it enters the receiver. In the 5000-horse-power engine 
at the Waterside Station in New York it was shown that both jackets 
and reheaters, either together or alone, were practically valueless 
throughout the working range of load. {Power, July, 1904, p. 424.) 
Many similar cases may be cited which show no gain in economy with 
the use of the reheaters. On the other hand, with properly propor- 
tioned reheaters, the gain may be considerable and particularly with 
superheated steam. Practically all European engines operating with 
highly superheated steam are equipped with receiver-reheaters. In all 
cases the reheater effects a great reduction in the condensation in the 
low-pressure cylinders, but the resulting gain, considering the conden- 
sation in the reheater coils, may be little if any. 

In triple expansion pumping engines receiver-reheaters are found to 
effect an appreciable gain in economy, and practically all such engines 
are equipped with them. In electric traction work or where the load 
is a widely fluctuating one the reheater has been virtually abandoned. 
Apart from the consideration of fuel economy, all tests show a marked 
increase in the indicated power of the low-pressure cylinder (5 to 15 
per cent), and to that extent it increases the capacity of the entire 
engine. (G. H. Barrus, Power, September, 1903, p. 516.) 

Receivers: Trans. A.S.M.E., 1-174, 178, 9-549; Ewing, Steam Engine, p. 222; 
Holmes, Steam Engine, p. 456; Spangler, Steam Engineering, p. 249; Whitham, 
Steam Engine, p. 395; Power, June, 1896, p. 20; ibid., Nov., 1905, p. 684. 

Receiver-Reheaters: Trans. A.S.M.E., 1-178, 17-509, 25-443; Power, Sept., 1903, 
p. 516; Am. Elecn., Oct., 1902, p. 480; ibid., July, 1904, p. 328; Eng. Rec., May 
9, 1903, p. 496; Engineering, Aug. 8, 1902, p. 197; ibid., Aug. 21, 1902, p. 125; Eng. 
Rec, May 28, 1904, p. 690. 

170. Jackets. — If the walls of the cylinder are made double and 
the space between is filled with live steam under boiler pressure, the 



STEAM ENGINES 289 

cylinder is said to be steam jacketed. The function of the jacket is 
to reduce initial condensation by maintaining the temperature of the 
internal walls as nearly as possible equal to that of the entering steam. 
The heat given up by the jacket steam, and the resulting condensa- 
tion, is usually a smaller loss than would otherwise result from cylinder 
condensation. However, tests of numerous engines with and without 
steam jackets do not agree as to the conditions under which their 
use is profitable, the apparent gain ranging from zero to 30 per cent. 
According to Peabody, a saving of 5 to 10 per cent may be made by 
jacketing simple and compound condensing engines, and a saving of 
10 to 15 per cent by jacketing triple expansion engines of 300 horse 
power and under. On large engines of 1000 horse power or more the 
gain, if any, is very small. (Peabody, "Thermodynamics," p. 400.) 

Other things being equal, the smaller the cylinder and the lower the 
piston speed the greater is the value of the jacket. Experiments 
show no advantage in increasing the jacket pressure more than a few 
pounds above that of the initial steam in the cylinder, and it is usual 
to reduce the pressure in the jackets of the second and succeeding 
cylinders of multi-expansion engines. (Ripper, " Steam Engine," p. 170.) 

To be effective, jackets should be well drained, kept full of live steam, 
and the water of condensation returned directly to the boiler. 

Pumping engines and other slow-speed engines running at practi- 
cally constant load are generally jacketed, but in street-railway work 
and in the majority of manufacturing plants carrying fluctuating load, 
jackets are not considered advantageous. 

Whatever may be the actual economy due to jacketing, there is no 
question but that the jacket greatly influences the action of the steam 
in the cylinders, and whether beneficially or not depends upon the 
design and construction of the engine. Unless otherwise specified, 
manufacturers usually build their engines without jackets. 

Steam Jackets: Trans. A.S.M.E., 1-175, 190, 2-198, 9-554, 11-141, 149, 328, 
1038, 12-873, 462, 13-176, 14-1356, 15-779, 137; Golding, 6<f> Diagram, p. 39; 
Hutton, Heat Engines; Perry, Steam Engine, p. 369; Rankine, Steam Engine, 
p. 395; Reeve, Thermodynamics, p. 29; Ripper, Steam Engine, p. 166; Thurston, 
Manual of the Steam Engine, 1-598, 622; Peabody, Thermodynamics, p. 322; Heck, 
Steam Engine, p. 123; Am. Mach., Jan. 30, 1896, p. 126; Engr., Lond., April 21, 1905, 
p. 401 ; Engineering, Jan. 30, 1905, p. 829; Eng. Rec, April 16, 1898, p. 423; Power, 
Feb., 1898, p. 17; ibid., Feb., 1899, p. 9; Eng. Mag., June, 1898, p. 479, June, 1899, 
p. 496, Aug., 1905, p. 755. 

Increasing Rotative Speed: See High-Speed vs. Low-Speed Engines, paragraph 172. 

Compounding : See Compound Engines, paragraph 177. 

Reducing Back Pressure : See Influence of Condensing, paragraph 179. 

Superheating : See paragraph 181. 



290 STEAM POWER PLANT ENGINEERING 

171. Single- and Double-Acting Engines. — When steam pressure is 
exerted on only one end of the piston the engine is said to be single 
acting, and when exerted alternately on one side and the other is said 
to be double acting. For high speed, minimum wear and tear, and 
comparatively cheap construction the single-acting engine offers some 
advantages. The Westinghouse Standard and the Willans Central- 
Valve engines are typical of this class. Silent running at high speed 
is possible because the pressure on the crank pin is not reversed. The 
output is only half that of a double-acting engine of the same size and 
speed, but a much higher rotative speed is permissible, which some- 
what offsets this disadvantage. Single-acting engines have been operated 
successfully with speed as high as 1000 r. p.m., while double-acting 
engines seldom exceed 350 r.p.m. and that only for strokes less than 
12 inches. 

Single-Acting Engines. — Comparison between Different Types of Engines ; Trans. 
A.S.M.E., 2-294. Economy of Single-Acting Expansion Engines: ibid., 3-252. Single- 
Acting Compound ; ibid., 12-275. Steam Distribution in Compound : ibid., 13-557; 
Ewing, Steam Engine, p. 371 ; Rankine, Steam Engine, p. 478 ; Hutton, Power Plants, 
p. 77. 

Double-Acting Engines: Hutton, Power Plants, p. 73; Rankine, Steam Engine, 
p. 50; Ewing, Steam Engine, p. 20; Le Van, Steam Engine, p. 240. 

173. High- and Low-Speed Engines. — High rotative speed does not 
necessarily mean high piston speed. An 8 x 10 engine running at 
300 r.p.m. has a piston speed of only 500 feet per minute, whereas a 
36 x 72 Corliss running at 60 r.p.m. has a piston speed of 720 feet per 
minute. The classification " high speed " and " low speed " refers to 
rotative speed only, the former above and the latter below say 150 
r.p.m. 

On account of the reduction of thermodynamic wastes, a high-speed 
engine should give theoretically a higher efficiency than the same engine 
at a lower speed, all other conditions being the same. The effect of 
speed upon economy is decidedly marked in engines and pumps taking 
steam full stroke. For example, tests of a 12 x 7J x 12 simplex 
direct-acting steam pump at Armour Institute of Technology showed 
a steam consumption of 300 pounds per I.H.P. hour at 10 strokes per 
minute, and only 99 pounds at 100 strokes per minute. (See Figs. 274 
and 275.) 

Tests of engines using steam expansively, however, do not furnish 
conclusive evidence on this point, some showing a decided gain (Pea- 
body, " Thermodynamics," p. 425), others little or no gain (Barrus, 
" Engine Tests," p. 260). For example, a small Willans engine showed 
an increase in economy of 20 per cent in increasing the rotative speed 



STEAM ENGINES 291 

from 111 to 408r.p.m. (Peabody, "Thermodynamics," p. 402), whereas 
the compound locomotives at the Louisiana Purchase Exposition showed 
a loss in economy for the higher speeds (Publication by the Penn- 
sylvania Railroad Company). On the other hand, a comparison of the 
performances of high- and low-speed Corliss engines shows little differ- 
ence in economy, and a general comparison between high- and low-speed 
engines furnishes little information, since nearly all high-speed engines 
are of a different class from the low-speed ones. High-speed engines are 
comparatively small in size, require larger clearance volume, and are 
usually fitted with a single valve. Rotative speed is limited by design, 
material, workmanship, and cost of subsequent maintenance. Speeds 
of 400 r.p.m. and more are not unusual with single-acting engines, 
whereas 300 r.p.m. is about the limit for double-acting machines with 
strokes over 12 inches in length. A comparison of tests of high-speed 
and low-speed engines in this country, irrespective of design and con- 
struction, shows the former to be less economical than the latter in 
most cases. In Europe high-speed engines are developed to a high 
degree of efficiency, and their performances are comparable with the 
best grade of low-speed engines. 

High-speed engines as a class have the advantage of being more 
compact for a given power, are simple in construction and relatively 
low in first cost; on the other hand, they are subject to comparatively 
rapid depreciation, excessive vibration, and are less economical in 
steam consumption. 

High- and Low-Speed Engines. — Effect of Speed on Condensation : Peabody, 
Thermodynamics, p. 424. Effect of Speed on Economy: Barrus, Engine Tests, 
p. 257; Trans. A.S.M.E., 7-397, 2-198. Limitation of Speed : Trans. A.S.M.E., 
14-806. Effect of Speed on Economy: Ripper, Steam Engine, p. 317. 

173-4. High-Speed Single-Valve Simple Engines. — This style of 
engine is made in sizes varying from 10 to 500 horse power. The 
cylinder dimensions vary from 4x5 to 24 x 24 and the rotative speed 
from 300 to 175 r.p.m. 

When ground is limited or costly and exhaust steam is necessary for 
heating or manufacturing purposes, the high-speed non-condensing 
engine is most suitable for horse powers of 200 or less, being 
compact, simple in construction and operation, and low in first cost. 
For sizes larger than this the compound engine would probably 
prove a better investment, except in cases where fuel is very cheap 
or large quantities of exhaust steam are to be used for manufacturing 
purposes. 

Small high-speed engines are seldom operated condensing, since the 



292 



STEAM POWER PLANT ENGINEERING 



gain due to reduction of back pressure is more than offset by the extra 
cost of the condenser and appurtenances. 

Engines are ordinarily rated at about 75 per cent of their maximum 
output. For example, a 12 x 12 non-condensing engine ^ running at 
300 r.p.m., with initial steam pressure of 80 pounds gauge, is nor- 
mally rated at 70 horse power, though it is capable of developing 90 
horse power at the same speed. 

The steam consumption of high-speed single-valve non-condensing 
engines at full load ranges from 27 to 50 pounds per indicated horse- 
power hour, depending upon the size of the unit and the conditions of 
operation. An average for good practice is not far from 30 pounds. 
With superheated steam a steam consumption as low as 18 pounds 
per horse-power hour has been recorded. 

Table 39 gives the steam consumption of a number of single-valve 
high-speed engines running condensing and non-condensing, and 



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Fig. 148. Typical Economy Curves of High-Speed, Single-Valve, Non-Condensing 
Engines. Saturated Steam. 

Fig. 148 shows some of the results for different loads. The steam 
consumption is fairly constant from 50 per cent of the rated load to 25 
per cent overload, but for earlier loads the economy drops off rapidly. 
The desirability of operating the engine near its rated load is at once 
apparent. The curves show a marked economy in favor of the larger 
cylinders, but the engines are not of the same make, and the conditions 
of operation are somewhat different. 



STEAM ENGINES 



293 




Fig. 148a. Assembly of Valve Gear; Typical Corliss Engine. 



Steam pipe 
| ">f Steam Flange 
.Corliss Steam Valve I I Throttle Valve 

Vanished Sheet Steel ^^_V<planrebed Steel Lagging 
Lagging / r q\ \ \ Heat IngnlatillBr FiUing 

Corliss Steam Valve Chamber 
Front Cylinder Head 
/ Front Cylinder Head Studs 



Back Cylinder Head 
Back Cylinder Head Studs 

Back Cyl. Head Bonnet - 



Corliss Exhaust Valve" 




^^Piston Rod Gland Studs 
:JV/^ /Piston Rod Gland 



-Piston Rod Packing 



Exhaust Chest 



Corliss Exhaust Valve 
lanished Sheet Steel Lagging 

Exhaust Flange x Heat Insulating Eilling 

Exhaust Opening 
Exhaust Pipe 



Fig. 148b. Section Through Cylinder; Typical Corliss Engine. 



294 



STEAM POWER PLANT ENGINEERING 




STEAM ENGINES 295 

The most economical cut-off for a simple engine is about one-third 
to one-fourth stroke when running non-condensing, and about one- 
sixth when condensing. 

An excellent performance for a small high-speed single-valve simple 
engine using saturated steam is shown by the Reeves piston valve 
engine (No. 6, Table 39) in a test made by Professor R. C. Carpenter. 
When running non-condensing with initial steam pressure of 114 pounds 
gauge the lowest steam consumption was 28 pounds per I.H.P. hour 
and 31.8 pounds per B.H.P. hour. Referred to the heat-unit basis, 
this gives 470 B.T.U. per I.H.P. per minute, non-condensing, and 450 
B.T.U. per I.H.P. per minute, condensing. 

A test of a 12 x 12 simple slide-valve Buffalo engine (No. 10, Table 
39) by Professor Reeves gave 30.6 pounds of steam per I.H.P. hour, 
initial gauge pressure 79.3 pounds. As far as the weight of steam and 
B.T.U. per I.H.P. per minute are concerned, the Reeves engine shows 
the better economy, but referred to the performance of the ideal engine 
the Buffalo engine is the more nearly perfect of the two. The latter con- 
sumed 510 B.T.U. per I.H.P. per minute, and a perfect engine working 
through the same range in temperature would require 324 B.T.U. per 
I.H.P. per minute; hence the efficiency ratio or the degree of perfection 
of the latter is 324 -r- 510 = 63.5 per cent as against 57.5 per cent for 
the former. 

Locomotive No. 1499 of the Pennsylvania system holds the record 
for economy in steam consumption for a single-valve non-condensing 
engine (No. 9, Table 39). The steam consumption is 23.4 pounds per 
I.H.P. hour (initial gauge pressure 196 pounds per square inch), which 
corresponds to a heat consumption of 398 B.T.U. per I.H.P. per minute. 
So far as the writer knows, this is the best recorded performance for a 
single- valve non-condensing engine using saturated steam. 

The performance of the Ames engine (No. 7, Table 39) is one of the 
best recorded for so low an initial pressure. 

A small single-acting Willans engine (No. 1, Table 39) holds an 
exceptional record for economy for a very small single-valve high-speed 
engine, having given a steam consumption of 26 pounds per I.H.P. 
hour, non condensing, initial gauge pressure 122 pounds, corresponding 
to a heat consumption of 436 B.T.U. per I.H.P. per minute. All of 
these performances are at the best rating of the engines. For a con- 
tinually changing load, as in electric -lighting service, the average steam 
consumption is considerably greater than that at full load and depends 
upon the "load factor" (the ratio of the actual to the rated load). 
This is clearly shown in the curves of the steam consumption, Figs. 148 
and 143. 



296 



STEAM POWER PLANT ENGINEERING 







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STEAM ENGINES 



297 



The performances given in Table 39 are exceptional. It is not ad- 
visable to count on a better steam consumption for this type of engine 
than 30 to 35 pounds of steam per I.H.P. hour. 















































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Fig. 149. 

Fig. 149 shows the effects of condensing on a typical single-valve 
high-speed engine. The gain in fuel economy may be only an apparent 
one, since the steam consumption of the condensing apparatus should 
be rightfully charged to the engine. 

When used in connection with heating plants or manufacturing plants 
requiring large quantities of exhaust steam the thermal efficiency is 
very high and may reach 85 per cent as against 22 per cent for the best 
compound condensing engine. In general when the requirements for 
exhaust steam are in excess of the steam consumption of a simple non- 
condensing engine a high-grade economical engine is without purpose. 

175. High-Speed Multi- Valve Engines. — The steam distribution in 
a single-valve engine may give good economy for a very small range in 
load but be far from satisfactory for a wide range. This must neces- 
sarily be so since admission, cut-off, release, and compression are all 
functions of one valve, and any change in one results in a change of the 
others. To obviate the limitations of the single valve, many builders 
design engines with two or more valves. With a two -valve engine cut- 
off is independent of the other events, and with four valves all events 
are independently adjustable. In addition to the flexibility of the valve 
gear, the chief feature of the four-valve engines lies in the reduction of 
clearance volume which is made possible by placing the valves directly 



298 



STEAM POWER PLANT ENGINEERING 



over the ports. The valves may be of the common slide-valve or rotary 
type. As a class, four-valve engines are more economical than those 
having a less number of valves. The advantages and disadvantages of 
the four-valve over the single-valve engines may be tabulated as below. 
Advantages. Disadvantages. 

1. Better steam distribution. 1. Increased number of parts. 

2. Better regulation. 2. Increased first cost. 

3. Reduced clearance volume. 3. Requires greater attention. 

4. Less valve leakage. 

5. Better economy. 

The steam consumption of a high-speed four-valve non-condensing 
engine varies from 22 to 35 pounds of saturated steam per horse-power 
hour, with an average not far from 27 pounds. With superheated steam 
the steam consumption may run as low as 18 pounds per horse-power hour. 

An exceptional performance for a simple unjacketed high-speed four- 
valve engine is that of Engine No. 17, Table 39. With initial gauge 
pressure of 125 pounds the steam consumption is 22.24 pounds per 
I.H.P. hour, corresponding to a heat consumption of 374 B.T.U. per 
I.H.P. per minute. 

100 



90 



o< 60 



50 



40 















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Comparative Economy 












of a 












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( B) Pour Valve High Speed 
Non Condensing Engine 












15 x 14 Reeves Simple (A) 












16 x 16 Memming Simple ( B ) 




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10 20 30 40 



50 60 70 80 90 
Per Cent of Rated Load 

Fig. 150. 



100 110 130 130 140 



Fig. 150 gives a comparison between a single- valve and a four- valve 
high-speed engine, and though the engines differ slightly in size, the 



STEAM ENGINES 299 

conditions of operation were comparable and the marked gain in 
economy of the latter over the former is apparent. Both perform- 
ances are exceptional, and a 10 to 15 per cent greater steam consump- 
tion may be expected in average good practice. 

As a general rule single- valve simple engines do not exceed 500 horse 
power in size for stationary work, whereas 1000 horse power is not an 
uncommon size for the multi-valve type. 

High-Speed Engines, General Description: Trans. A.S.M.E., 2-75; ibid., 17-117; 
Engr. U.S., Jan. 15, 1903; Engr., Lond., April 15, 1904, p. 379, April 29, p. 433, 
May 13, p. 478, May 20, p. 529. Proportions of High-Speed Engines: Trans. 
A.S.M.E., 8-191; Klein, High-Speed Engines; Hutton, Power Plants, p. 70, Thurs- 
ton, Stationary Steam Engines. 

Tests of Simple High-Speed Engines : Am. Elecn., April, 1901, p. 197, Dec, 1903. 
p. 581; Trans. A.S.M.E., 11-723, 18-795; Elec. World, May 20, 1903, p. '897, Sept. 
10, 1904, p. 404, Oct. 1, 1904, p. 587, Feb. 17, 1906, p. 369; Engr. U.S., June 1, 1903, 
p. 416, Nov. 1, 1904, p. 758; Engineering, July 22, 1898, p. 116; Eng. News, Dec. 3, 
1903, p. 493; Eng. Rec, July 6, 1901, p. 225; Machinery, May, 1903, p. 481; Power, 
Jan., 1904, p. 44, Nov., 1904, p. 651, Jan., 1905, p. 56; Stevens Indicator, Jan., 
1900, p. 9; St. Ry. Jour., Oct., 1904, p. 673; Technology Quarterly, Sept., 1899, 
p. 255. 

176. Medium and Low-Speed Multi- Valve Engines. — A comparison 
of tests of high and low-speed single-valve engines irrespective of design 
and construction shows the former as a class to be less economical than 
the latter. With four-valve engines there is no such disparity, and the 
high-speed type has shown just as good economy as the slow-speed class. 
For example, Engine No. 17, Table 39, with Corliss valves and a speed of 
210 r.p.m., gives practically the same economy as Corliss engine No. 15 
operating at 62 r.p.m. By far the greater number of simple multi- 
valve slow-speed simple engines are of the Corliss type. They range 
in size from 50 to 3000 horse power, with cylinders varying from 12 x 30 
to 48 x 72. The smaller sizes with trip-valve gear run at 90 to 100 
r.p.m., and the larger at 50 to 75 r.p.m. Without the trip gear, speeds 
of 150 r.p.m. are not uncommon, but at this speed they are usually 
classified as high-speed engines. 

Table 39 gives the steam consumption, condensing and non-con- 
densing, of a number of four- valve slow-speed simple engines. 

Engine No. 15 shows an unusual performance for a simple Corliss 
engine operating both condensing and non-condensing. With initial 
gauge pressure of 103.5 pounds, the minimum steam consumption is 
21.5 pounds per I.H.P. hour for the non-condensing run and 16.5 pounds 
for the condensing run. This corresponds to 358 B.T.U. per I.H.P. 
per minute, non-condensing, and 302 B.T.U. per I.H.P. per minute, 



300 STEAM POWER PLANT ENGINEERING 

condensing. The efficiency ratios are 78.0 per cent and 53.2 per cent 
respectively. The cylinder was jacketed. 

Attention is also called to the record of engine No. 22, Table 39, 
which is of the Sulzer type with four balanced poppet valves, heads 
and cylinder barrel jacketed. With 79 pounds initial pressure and a 
vacuum of 1.36 pounds absolute, the steam consumption is 15 pounds 
per I.H.P. hour, corresponding to a heat consumption of 275 B.T.U. 
per I.H.P. per minute. 

177. Compound Engines. — Compound engines may be divided 
into three classes, tandem, cross compound, and duplex. In the 
tandem the two cylinders are end to end, in the cross compound side 
by side, and in the duplex one above the other. The tandem and 
duplex compounds have the advantage of (1) compactness for a 
given power, (2) less complication and fewer parts, and (3) low 
first cost. The crank effort is more variable than in the cross com- 
pound. In very large engines the low-pressure stage is generally 
divided between two cylinders of equivalent size to avoid an excess- 
ively large single cylinder and to distribute the crank effort. High- 
speed non-condensing compounds are ordinarily of the tandem type 
and are finding much favor in isolated station work, as in the 
power plants of tall office buildings where ground space is limited, 
though the duplex compound is sometimes used. The vertical or 
horizontal cross compound is generally installed in street-railway 
plants. 

Cylinder ratios for high-speed single-valve compound engines vary 
from about 1 to 2\ with 100 pounds pressure to about 1 to 3 with a 
pressure of 150 pounds, and for slow-speed condensing engines from 1 
to 3 with 125 pounds pressure to about 1 to 4 with a pressure of 175 
pounds. G. I. Rockwood recommends a ratio as high as 7 to 1, and a 
number of engines designed along this line have shown exceptional 
economy. A cross compound Corliss engine at the Atlantic Mills, 
Providence, R.I., with cylinders 16 and 40x48 (ratio 6.128 to 1) 
gave the low steam consumption of 11.2 pounds of steam per I.H.P. 
hour, corresponding to a heat consumption of 222 B.T.U. per I.H.P. 
per minute. The 5500-horse-power engines of the New York Edison 
Company have a cylinder ratio of 6 to 1. The great majority of corn- 
pound engines, however, have cylinder ratios of 4 to 1 or less. The 
8000-horse-power engines of the Interborough Rapid Transit system 
have a ratio of 4 to 1, and the 4000-horse-power units of the Metro- 
politan Elevated Company, New York, a ratio of 3.5 to 1. 



STEAM ENGINES 



301 




Fig. 150a. 3500 K.W. Vertical Cross-Compound Corliss Engine as Installed at the 
Power House of the Twin City Rapid Transit Co., Minneapolis, Minn. 



302 



STEAM POWER PLANT ENGINEERING 







Fig. 150b. 7500 K.W. Vertical-Horizontal Double Compound Engine as Installed at the 
59th Street Station of the Interborough. (Manhattan Type.) 



STEAM ENGINES 



303 



The respective advantages and disadvantages of compounding may 
be tabulated as follows: 



Advantages. 

1. Permits high range of expansion. 

2. Decreased cylinder condensation. 

3. Decreased clearance and leakage 



4. Equalized crank effort. 

5. Increased economy in steam 

consumption. 



Disadvantages. 

1. Increased first cost due to multi- 

plication of parts. 

2. Increased bulk. 

3. Increased complexity. 

4. Increased wear and tear. 

5. Increased radiation loss. 



The ratio of expansion for a multi-expansion engine is usually 
taken to be the product of the ratio of the volume of large to small 
cylinder divided by the fraction of the stroke at cut-off in the high- 
pressure cylinder. For example, a compound engine with cylinders 
24, 48 x 48 cutting off at ^ in the high-pressure cylinder has a nominal 
ratio of expansion of 4 -j- | = 12. The number of expansions at 
rated load in compound condensing engines varies widely, ranging 
from 10 to 33, with an average not far from 16. 

The steam consumption shown by tests of a number of compound 
engines using saturated steam, condensing and non-condensing, is 
given in Table 40. For tests with superheated steam see Table 43. 





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\ 




































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Belative Economy 

of a 

Simple and Compound 

Non-Condensing High Speed 

Engine 












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30 30 40 50 60 70 80 90 100 120 140 

Developed Horse Power (B. H. P.) 

Fig. 151. 



160 



180 



Fig. 151 shows the relative economy under comparable conditions 
of a high-speed simple and a high-speed compound engine, both run- 
ning non-condensing and using saturated steam. The advantage of 
the compound at full load and overload is very marked, though its 



304 



STEAM POWER PLANT ENGINEERING 



economy drops off rapidly at light loads and may be less than that of 
the simple engine. 

Fig. 152 shows the relative economy of two compound Corliss engines 
running condensing and non-condensing, both using saturated steam. 







































1 








































A 21,41 x 30 Compound 
B 20,40 x 42 Compound 










































20 


















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300 400 500 600 700 800 900 1000 1100 1200 1300 1400 
Indicated Horse Power 

Fig. 152. 



It should be borne in mind that the object of compounding is to 
permit the advantageous use of high pressures and large ratios of 
expansion. Under proper conditions compounding may increase the 
economy at rated load about 20 per cent for non-condensing engines 
and 30 per cent for condensing engines. 

An exceptional performance of a single-valve high-speed non-con- 
densing compound engine is that of engine No. 20, Table 40. With 
initial gauge pressure of 128 pounds the steam consumption is 22.3 
pounds per I.H.P. hour, corresponding to a heat consumption of 376 
B.T.U. per I.H.P. per minute. 

One of the best performances of a multi- valve high-speed compound 
non-condensing engine is that of engine No. 14, Table 40. With 
initial pressure of 175 pounds gauge the steam consumption at full 
load is 17.17 pounds per I.H.P. hour, corresponding to a heat con- 
sumption of 291 B.T.U. per I.H.P. per minute. 

The 8000-horse-power vertical cross compound Corliss engines of 
the Interborough Rapid Transit system (No. 6, Table 40), probably 
hold the record for economy for compound engines without jackets 
and reheaters, using saturated steam. With initial pressure of 175 
pounds gauge and absolute back pressure of 2.2 pounds, the steam 
consumption is 11.96 pounds per I.H.P. hour, corresponding to a heat 



STEAM ENGINES 



305 



consumption of 220 B.T.U. per I.H.P. per minute. In estimating aver- 
age practice it would be safe to add 10 per cent or 20 per cent to the 
steam consumptions given in Table 40. 



Lb 
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Fig. 153. Economy Test of the 5500-Horse-Power Three-Cylinder Compound Engine 
and Generator at the Waterside Station of the New York Edison Co. 

Fig. 153 illustrates the performance of the 5500-horse-power three- 
cylinder compound engine at the Waterside Station of the New York 
Edison Company. The best economy is 11.93 pounds of steam per 
I.H.P. hour, corresponding to a heat consumption of 221 B.T.U. per 
I.H.P. per minute. 

Compound Engines. — Best Load for Compound Engine : Trans. A.S.M.E., 18-674. 
Cylinder Proportions for Compound and Triple Expansion Engines: Trans. A.S.M.E., 
21-1002, 16-762; Engr. U.S., Sept. 1, 1906, p. 586; Eng. News, March 2, 1899, p. 137; 
Eng. Rec, Jan. 7, 1899, p. 122; Power, June, 1904, p. 47. Laws of the Average 
Simple vs. Compound Engines under Variable Load: Am. Mach., Sept. 27, 1900, p. 927; 
Non-condensing Compound Engine for Office Buildings: Eng. Rec, June 18, 1898, 
p. 45. Economical Use of Steam in Non-condensing Engines: Eng. Mag., May, 1898, 
p. 213, July, 1898, p. 603. 

Compound Engine Tests: Trans. A.S.M.E., 24-1274, 25-264; Engr. Lond., 99-546; 
Eng. News, Jan. 11, 1906, p. 44; Eng. Rec, April 16, 1898, p. 431, June 4, 1898, 
p. 1; Nov. 18, 1899, p. 579; Sibley Jour., May, 1901, p. 346; St. Ry. Jour., 27-41. 

178, Triple and Quadruple Engines. — Triple and quadruple expan- 
sion engines are in general use where the load is practically constant, 
as in marine and pumping-station practice, but have been abandoned 
in street-railway work and in plants where the load fluctuates widely, 



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STEAM ENGINES 307 

in favor of the two or three-cylinder compound. The best economy on 
a heat-unit basis ever recorded for an engine using saturated steam was 
that of the Nordburg quadruple pumping engine at Wildwood, Pa., 
which gave a consumption of 12.26 pounds per I.H.P. hour and a heat 
consumption of 186 B.T.U. per I.H.P. per minute reckoned above the 
feed-water temperature.* The Allis triple expansion pumping engine 
at Chestnut Hill holds the record for saturated steam consumption, 
10 pounds per I.H.P. hour, and its exceptional performance of one 
developed horse power per 1.09 pounds of coal has, perhaps, never 
been excelled. An inverted vertical marine cross compound engine, 21 
and 36 x 36, built by Cole, Marchent & Morley, Bradford, England, 
holds the record for superheated steam consumption, 8.58 pounds per 
I.H.P. hour. (Table 47.) On the heat basis (192 B.T.U. per I.H.P. per 
minute), however, it does not equal the performance of the Nordburg 
engine. The above efficiencies have been exceeded by the binary 
vapor engine at Berlin; but this belongs in a class by itself and should 
hardly be compared with the ordinary form of steam engine. (See 
paragraph 182.) 

Triple Expansion Engines. — Cylinder Proportions for Triple Expansion Engines : 
Trans. A.S.M.E., 21-1002, 10-576. Economy of Triple Expansion Engines : Trans. 
A.S.M.E., 8-496. 

179. Effects of Condensing. — The effect of the condenser upon the 
power and economy of engines is indicated in Table 41. The curves 
in Figs. 154 and 155 were plotted from tests made by Professor R. L. 
Weighton on a 7, 10^, 15J x 18 triple expansion engine at Durham 
College of Science, Newcastle-on-Tyne. The straight line shows how 
the mean effective pressure would vary with the degree of vacuum if 
the power increased directly with the reduction in back pressure. 
The curved line shows the actual M.E.P., which increases almost along 
the theoretical line up to a 10-inch vacuum, from which point on the 
increase is less marked. At 26 inches the actual M.E.P. reaches an 
apparent maximum. These figures are not applicable to all engines but 
give a good idea of the limitation of the vacuum with the reciprocating 
engine. The gain in steam consumption due to the condenser does 
not indicate a corresponding gain in heat consumption. For example, 
Engine No. 2, Table 41, shows an apparent gain in steam consumption, 
due to condensing, of 12.5 per cent, the temperature of the feed water 
returned to the boiler being 120 degrees F. With a suitable heater 
the exhaust of the non-condensing engine would be capable of heating 

* Replaced in 1905 by a Riedler pumping engine on account of high maintenance 
cost. 



308 



STEAM POWER PLANT ENGINEERING 



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STEAM ENGINES 



309 



the feed water to 210 degrees F. The non-condensing engine should 
therefore be credited with 210 - 120 or 90 heat units per pound of 
steam used, or, in round numbers, 9 per cent. The difference between 
12.5 per cent and 9 per cent, or 3.5 per cent, represents the net gain in 
favor of condensing provided the power necessary to create the vacuum 
is ignored. Actually the steam consumption of the condenser pumps 
might be equal to or greater than 3.5 per cent of the steam generated 
and the net gain becomes zero or even negative. Referring to Fig. 
155, plotted from tests of the 7, 10i, 15J x 18 triple expansion engine 
mentioned above, the solid lines show the feed-water consumption per 
I.H.P. hour and the broken line the heat units consumed per brake 



t a 



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Fig. 156. Performance of 5500 H. P. Engine at Waterside Station of New York 

Edison Company. 

horse power per minute measured above the hot-well temperature. 
The engine efficiency, based upon the water consumption, increases as 
the vacuum increases, reaching a maximum between 26 and 28 inches, 
whereas the heat-unit curve gives the maximum between 20 and 21 
inches. Between 22 and 28 inches the heat-unit curve shows a rapid 
falling off in economy. Tests of the 5500-horse-power engine at the 
New York Edison Company's Waterside Station showed that increasing 
the vacuum from 25.3 to 27.3 inches decreased the water rate only 
0.06 pound per I.H.P. {Power, July, 1904, p. 424.) The results are 
illustrated in Fig. 156. In most cases, and particularly with large 



310 



STEAM POWER PLANT ENGINEERING 



compound engines, the net gain due to condensing is considerable, but 
the feed-water temperatures and power consumed by the auxiliaries 
should be taken into account.* Fig. 149 shows the effect of vacuum 
on the steam consumption of a small high-speed simple engine, and 
Fig. 152 of a cross compound Corliss. (See also paragraph 210.) 



TABLE 41. 

EXAMPLES OF THE EFFECT OF CONDENSING ON THE ECONOMY OF 
RECIPROCATING ENGINES. 





Non-Condensing. 


Condensing. 


Increase Due 
to Condensing. 


O) CD 
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Pounds per 
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a 

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1 

2 
3 
4 
5 
6 
7 
8 
9 

10 
11 


147 
148 
126 

67.6 
103.8 
114 

96 
118 

75.9 

62.5 
186.7 


54.7 
540 

83 
209 
177.5 
160 
120 
267 
310 
451 

40.4 


19.2 

19.3 

23.8 

28.9 

22.1 

31 

23.9 

23.24 

25.6 

30.1 

18.7 


149 
147 
130 

67 
103.8 
114 

96 
119 

79 

63.6 
184.6 


1.6 

4 

7.4 

4.5 

1.2 

"4' 

4.2 
6.4 
7.8 
1.6 


83.4 

116 
213 
155 
168 
145 
276.9 
336 
444 
29.8 


14.8 

16.9 

19.1 

22 

16.5 

27 

19.4 

16 

20.5 

23 

12.7 


52.5 

* 

39.8 

1.9 

* 

2 

20.8 
3.7 

8.7 
* 
* 


25 

12.5 

19.7 

23.5 

25.1 

12.9 

18.8 

31 

19.9 

23.6 

32 



Cut-off changed for best economy. 

7, 10J, 15i x 18 triple ; Eng. News, Aug. 21, 1902, p. 127. 

17, 27 x 24 Westinghouse marine, non-condensing ; Power, August, 1903. 

1, 18 x 10 Buffalo tandem compound ; Elec. World, Sept. 10, 1904, p. 404. 

18 x 30 four-valve (slide) ; Engine Tests, Barrus, p. 88. 

21, 65 x 43.31 Corliss ; Peabody's Thermodynamics, p. 382. 

12 x 12 Reeves simple ; Elec. World, Oct. 1, 1904, p. 587. 

18 x 48 simple Corliss ; Peabody's Thermodynamics, p. 354. 

14, 28 x 24 two- valve (slide) ; Engine Tests, Barrus, p. 175. 

17 x 24 two-valve; Engine Tests, Barrus, p. 70. 

28 x 36 Corliss ; Engine Tests, Barrus, p. 97. 

Willans triple expansion central valve engine ; Peabody, Thermodynamics, p. 406. 



180. Throttling vs. Automatic Cut-Off. — The action of the gov- 
ernor in the throttling engine is shown by the superposed indicator 
cards (Fig. 157) taken between zero or friction load and maximum 
load. The effect of throttling is to reduce the pressure during admis- 
sion, but does not change the point of cut-off or other events of the 
stroke. The steam may be partially dried or even superheated by 
throttling, thus tending to reduce cylinder condensation. Initially 
dry saturated steam at a pressure of 125 pounds gauge would be super- 
* See Power, Feb. 23, 1909, p. 381. 



STEAM ENGINES 



311 



heated about 12 degrees in expanding through a throttle to 90 pounds, 
or if it contained initially 2 per cent moisture would be perfectly 
dried in expanding to 40 pounds. (See Table 42.) Friction through 
the valve also tends to dry the steam. Thus with very light loads 
the superheat may be decidedly appreciable. The possible gain due 




Fig. 157. Typical Indicator Cards. High-Speed Throttling Engine. 

to decreased cylinder condensation is to some extent offset by incom 
plete expansion. The best efficiency for a given load is realized b} 
a proper compromise between cut-off and initial pressure. Experi- 
ments made by Professor Denton (Trans. A.S.M.E., 2-150) on a 
17 x 30 non-condensing double-valve engine showed the most economical 
results with J cut-off for 90 pounds pressure, J cut-off for 60 pounds, 
and tVq for 30 pounds. The average throttling engine does not give 
close regulation, the governor usually lacking sensitiveness. Tests 
show the economy to be better than that of the automatic engine on 
light loads, and the crank effort more uniform. 




Fig. 158. Typical Indicator Cards. High-Speed Automatic Engine. 



The indicator cards shown in Fig. 158 were taken from a single- 
valve high-speed automatic engine operating between friction load and 
maximum load. The mean effective pressure is adjusted to suit the 
load by the automatic variation in the cut-off, the initial pressure 
remaining the same. Since the cut-off is controlled by the action of 
the governor on the single valve, all other events of the stroke are 



312 



STEAM POWER PLANT ENGINEERING 



likewise cnanged. With a four-valve engine the variation in cut-off 
does not affect the other events. 

The chief advantage of the automatic over the throttling engine lies 
in its sensitive regulation, and while, in general, it gives a lower 
steam consumption than the throttling engine, this is probably in 
most cases due to superior construction and not to the method of 
governing. 

TABLE 42. 

SHOWING THE INITIAL PER CENT OF MOISTURE THAT WILL BE EVAPORATED 
IN THROTTLING FROM A HIGHER TO A LOWER PRESSURE. 

Based on Marks' and Davis' Steam Tables. 



Final Pressures. 






I 


nitial Pressure, 


Absolute 










80 


85 


90 


95 


100 

0.45 
0.59 
0.74 
0.88 
1.06 
1.23 
1.43 
1.66 
1.91 
2.20 
2.53 
2.92 
3.40 
4.01 


105 


110 


115 


120 


80 




0.13 
0.26 
0.40 
0.55 
0.71 
0.89 
1.09 
1.32 
1.56 
1.85 
2.18 
2.56 
3.04 
3.65 


0.24 
0.37 
0.52 
0.66 
0.83 
1.01 
1.21 
1.44 
1.68 
1.97 
2.30 
2.69 
3.16 
3.78 


0.36 
0.49 
0.64 
0.78 
0.95 
1.13 
1.33 
1.56 
1.80 
2.10 
2.42 
2.82 
3.29 
3.90 


0.55 
0.70 
0.84 
0.99 
1.16 
1.34 
1.54 
1.76 
2.02 
2.31 
2.64 
3.03 
3.51 
4.13 


0.65 
0.79 
0.93 
1.08 
1.25 
1.44 
1.64 
1.86 
2.12 
2.41 
2.74 
3.13 
3.61 
4.23 


0.74 
0.88 
1.03 
1.18 
1.34 
1.53 
1.74 
1.96 
2.M 
2.51 
2.84 
3.23 
3.71 
4.33 


83 


75 


0.14 
0.28 
0.43 
0.59 
0.77 
0.97 
1.19 
1.44 
1.72 
2.05 
2.44 
2.90 
3.51 


0.97 


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1 12 


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60 

55 

50 

45 

40 

35 

30 

25 

20 


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1.44 
1.62 
1.82 
2.05 
2.30 
2.60 
2.93 
3.32 
3.80 


15 


4.43 



Final Pressures. 






Initial Pressure, Absolute 










125 


130 


135 


140 


145 


150 


155 


160 


165 


80 


0.91 
1.05 
1.19 
1.34 
1.52 
1.70 
1.90 
2.13 
2.39 
2.68 
3.01 
3.41 
3.88 
4.51 


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1.27 
1.43 
1.60 
1.78 
1.99 
2.21 
2.47 
2.77 
3.10 
3.49 
3.97 
4.60 


1.08 
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1.36 
1.51 
1.68 
1.86 
2.08 
2.30 
2.55 
2.85 
3.18 
3.58 
4.06 
4.70 


1.15 
1.28 
1.43 
1.59 
1.76 
1.94 
2.15 
2.38 
2.63 
2.93 
3.26 
3.66 
4.15 
4.78 


1.22 
1.36 
1.50 
1.66 
1.83 
2.02 
2.22 
2.45 
2.71 
3.01 
3.34 
3.74 
4.22 
4.86 


1.29 
1.43 
1.58 
1.73 
1.90 
2.09 
2.30 
2.52 
2.78 
3.08 
3.41 
3.81 
4.30 
4.94 


1.35 
1.49 
1.64 
1.79 
1.96 
2.15 
2.36 
2.59 
2.84 
3.14 
3.48 
3.88 
4.37 
5.01 


1.41 
1.55 
1.70 
1.85 
2.03 
2.21 
2.42 
2.65 
2.91 
3.21 
3.55 
3.96 
4.45 
5.09 


1.48 


75 


1.62 


70 


1.77 


65 


1.93 


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2.10 


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45 


2.29 
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2.73 


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2.99 


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3.29 


30 

25 

20 

15 


3.63 
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4.53 
5.17 



STEAM ENGINES 



313 



The following performances of a Belliss 250-horse-power high-speed 
condensing engine fitted with both automatic and throttling govern- 
ing devices give results decidedly in favor of the throttling engine. 
(Pro. Inst, of Mech. Engrs., 1897, p. 331.) 



Percentage of load .... 
Electrical horse power. 
Steam per I.H.P. hour. 



Automatic Cut-Off. 




Throttling. 


100 


62.5 


33 


25 


100 


62.5 


33 


213 


132 


77.8 


53 


213 


132 


77.8 


22.5 


22.9 


28.5 


34.3 


21 


21.7 


25.6 



25 
53 

28.4 



Some of the comparative advantages and disadvantages of the 
automatic and throttling engines are as follows: 



Automatic. 

1. Sensitiveness of regulation. 

2. Increased ratio of expansion. 

3. Low terminal pressures. 



Throttling. 



Advantages. 



1. Low first cost. 

2. Crank effort more uniform. 

3. Reduced cylinder condensation. 

4. Simplicity of regulating device. 



1. Increased cylinder condensation. 

2. Greater variation in crank effort. 

3. Complicated valve gear. 

4. Low economy at very early loads. 



1. Low ratio of expansion. 

2. High terminal pressure. 

3. Low initial pressure at early loads. 



181. Influence of Superheat. — (See also paragraph 103.) Table 43 
gives test results for several different types of engines employing super- 
heated steam. These figures may be compared with the perform- 
ances of engines using saturated steam as given in Tables 39 and 40. 
A decided gain in economy is shown in favor of superheat for single- 
cylinder engines. With compound engines the advantage is not so 
apparent, while triple expansion engines show the least gain. Tables 
44 to 46 show the effect of superheating on simple, compound, and 
triple expansion engines. (Proc. A.S.M.E., September, 1907.) As far 
as steam consumption is concerned, most engines show greater economy 
with superheated than with saturated steam, but the gain in thermal 
efficiency is not so marked, and when the economy is measured in 
dollars and cents per developed horse power, taking all things into 
consideration the gain is still further reduced and in many cases com- 
pletely neutralized. 



314 



STEAM POWER PLANT ENGINEERING 



Fig. 159 gives the results of a series of tests made on a number of 
Belliss & Morcom engines using superheated steam. (Pro. Inst, of 
Mech. Engrs., March, 1905, p. 302.) The engines were from 200 to 
1500 kilowatts capacity and were tested at full load. It is noticeable that 
the curves all converge to a single point and will meet at about 400 




100 



200 
Superheat, Deg. F. 



Fig. 159. Effect of Superheat on Steam Consumption. 



degrees F. The results show that if sufficient superheat is put into 
the steam all engines of whatever size are equally economical. Fig. 160 
shows the relationship between degree of superheat and the heat 
consumption at various loads for a 300-horse-power Belliss & Morcom 
high-speed triple expansion engine. (Pro. Inst. Mech. Engrs., March, 
1905, p. 303.) It will be noted that the variation in heat consump- 
tion at different percentages of load becomes less marked as the degree 
of superheat increases. With superheat of 350 degrees F. the heat 
consumption from \ load to full load is practically constant. . 



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U 



STEAM ENGINES 



319 




Fig. 159a 3000 H.P. Sulzer Engine Designed for Highly Superheated Steam. 



320 



STEAM POWER PLANT ENGINEERING 




3,. 

: to 

J a 

J : KJ 



d 



ft 



-^ J -d 



STEAM ENGINES 



321 



Table 47 gives the results of a test made on a 21, 36 x 36 inverted 
vertical marine engine. (Engr., Lond., June 2, 1905, p. 546.) 

TABLE 47. 

PERFORMANCE OF 21, 36 X 36 INVERTED VERTICAL MARINE CROSS 
COMPOUND ENGINE. 

{Engr., Lond., June 2, 1905, p. 546.) 



Pressure on boiler side of throttle valve, 
gauge 

Temperature of steam on boiler side of 
throttle, degrees F 

Degrees superheat of steam on boiler 
side of throttle, F 

Temperature of steam at admission, F. . . 

Degrees superheat of steam at admis- 
sion, F 

Vacuum, inches of mercury, absolute. . . 

I.H.P 

M.E.P. referred to L.P. cylinder 

Revolutions per minute 

Pounds of steam per I.H.P. hour 

B.T.U. per I.H.P. per minute 

See footnote * 

B.T.U. per I.H.P. per minute, perfect 
engine 

Thermodynamic efficiency 

Efficiency ratio 

Equivalent evaporation of saturated 
steam reckoned from hot well 

Temperature of hot well 



117.5 


117.5 


117.5 


117 


114.5 


743 


738 


749 


751 


732 


395 


390 


401 


403 


384 


601 


590 


569 


580 


558 


253 


242 


221 


232 


210 


3.4 


3 


2.22 


1.82 


1.82 


481.3 


461.1 


347.5 


333.5 


258 


26 


29.9 


18.8 


18 


14.4 


100.6 


100.7 


100.6 


100.7 


100.7 


9.09 


9.26 


8.88 


8.68 


8.74 


192.2 


201.7 


197.6 


194 


194 


187 


189 


181 


179 


179 


142.4 


142.5 


130.2 


126 


128.5 


21.4 


21 


21.4 


21.8 


21.8 


72 


72 


66 


65 


66 


10.63 


10.81 


10.38 


10.07 


10.12 


102 


101 


78 


71 


64 



114.5 

726 

378 
550 

202 

1.7 
145.52 

7.87 
100.7 

8.58 
192.1 
175 

128 
22 
67 

10.03 
70 



* B.T.U. per I.H.P. per minute based on latest (May, 1908) values for specific heat of super- 
heated steam. 



The relationship between the weight of steam consumed per I.H.P. 
hour and the equivalent heat consumption of a 250-horse-power tan- 
dem compound Van Den Kerchove engine is illustrated in Fig. 161. 

The performances of engines using superheated steam should be 
expressed in B.T.U. per I.H.P. per minute or the equivalent, as the 
steam consumption alone gives no idea of the true heat consumption. 

182. Binary- Vapor Engines. — A consideration of the Carnot or 
Rankine cycles shows that theoretically the efficiency of the steam 
engine may be increased by raising the temperature of the steam 
supplied or by lowering the temperature of the exhaust, that is to say, 
by increasing the range. Superheated steam development has prac- 
tically determined the upper limit, and economical practice indicates a 
vacuum of about 26 inches, corresponding to 126 degrees F., as the 
average lower limit for most efficient results from a commercial stand- 
point. 



322 



STEAM POWER PLANT ENGINEERING 

• 



20 




\\ 






































IS 


I 






















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40 60 

Per Cent of Bated Load. 



loo 



Fig. 160. Effect of Superheat on Steam Consumption. 



( 
































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245 


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Influence of Superheat on Economy 

(250 H.P. Tandem Comp'd Van Den) 

Kerchove Engine 












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190 











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Degrees of Superheat, Fahrenheit 
Fig. 161. 



STEAM ENGINES 



323 



In the binary-vapor engine the working range has been considerably- 
increased by substituting a highly volatile liquid, as sulphur dioxide, 
for the water which is ordinarily used as the cooling medium in the 
surface condenser. 

The S0 2 in condensing the exhaust steam is itself vaporized and the 
vapor, under a pressure of about 175 pounds per square inch, used 
expansively in a secondary reciprocating engine. The exhausted S0 2 
is discharged into a surface condenser in which it is liquefied by cooling 
water much the same as in refrigerating practice and used over and 







r 

! 


-To AicPump 




F 

SO Vaporizer and Steam 
Condenser 
l87#A1jsolutel48 F 


SOgVapor 


1 


1 




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S0 2 Tank 


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155 ° p ..!l 




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Cylinder 












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Cylinder 


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J A 






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Circulating 

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jsolute 

143.5 R.P.M. 































Fig. 162. Diagram of Binary- Vapor Engine. 



over again. Referring to Fig. 162, which illustrates diagrammatically 
a binary-vapor engine at the Royal Technical High School, Berlin: 
A, B, and C are the three steam cylinders of an ordinary triple expan- 
sion engine and D the S0 2 cylinder. All four cylinders drive a common 
crank shaft E. F is a high-pressure surface condenser which acts as 
a vaporizer for the S0 2 and a condenser for the steam. G is a surface 
condenser which serves to condense the S0 2 vapor. H is a liquid S0 2 
tank. The operation is as follows: Highly superheated steam enters 
the high-pressure steam cylinder at I and leaves the low-pressure 
cylinder at /, just as in any steam engine. The exhaust steam enters 
chamber F and is condensed by the liquid S0 2 passing through the 
coils. The condensed steam and entrained air are removed from the 
chamber by a suitable air pump. The steam in condensing gives up 
its latent heat to the liquid S0 2 and causes it to vaporize. The S0 2 
vapor passes from the coils in chamber F to the S0 2 engine D and 
performs work. The exhausted S0 2 vapor flows from cylinder D to 
chamber G, and is condensed by cooling water flowing through a series 



324 STEAM POWER PLANT ENGINEERING 

of tubes. The liquid S0 2 is collected in liquid tank H and thence is 
pumped into the coils in vaporizer F. The approximate temperatures 
and pressures at different points of the cycle are indicated on the 
diagram. 

A number of experiments made by Professor E. Josse in the labora- 
tory of the Royal Technical High School of Berlin on an experimental 
plant of about 200 horse power gave some remarkable results. A few 
of the tests made with highly superheated steam gave the following 
average figures : 

I.H.P. (steam end) 146.4 

Steam consumption per I.H.P. hour 12.8 

I.H.P. (S0 2 end) , 52.7 

Percentage of power of S0 2 engine 35.9 

Steam consumption per I.H.P. hour of combined engine 9.43 

When operating under the most satisfactory conditions a perform- 
ance of 8.36 pounds of steam per I.H.P. hour was recorded, correspond- 
ing to a heat consumption of 158.3 B.T.U. per minute, which is the 
best recorded performance to date (1907) in the history of steam- 
engine economy. 





























































































































































































4000 


























































































































3500 


























































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50 75 100 150 200 250 300 350 400 

Horse-power 

Fig. 163. Cost of Simple High-Speed Engines. 



S0 2 does not attack the metal surface of the engine unless combined 
with water, in which case sulphurous acid is formed. There is, how- 
ever, no danger from this cause, since the S0 2 being under greater 
pressure effectually prevents leakage of water into the S0 2 system. 



STEAM ENGINES 



325 



The S0 2 cylinder requires no other lubrication than the S0 2 itself, 
which is of a greasy nature. 





























































































































/ 


























































s 


y 




40000 






















































y 
























































y 










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30000 








































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i 


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15000 
















































































S 


» 








































10000 
















s 


/ 








































































































5000 
8500 






























































J* 


























































y * 

























































200 400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 
■SDrse-powner 

Fig. 164. Cost of High-Speed Compound Engines. 




100 200 300 400 500 600 700 800 900 1000 1100 1200 
Hoige-power 

Fig. 165. Cost of Low-Speed Engines, Simple and Compound. 



Properties of S0 2 : Trans. A.S.M.E., 25-181 . Binary-Vapor Engines : Jour. Frank. 
Inst., June, 1903; Elec. World and Engr., Aug. 10, 1901; U.S. Cons. Reports, 
No. 1139, Sept. 14, 1901; Engr. U.S., Aug. 1, 1903; Sib. Jour, of Eng., March, 1902. 



326 STEAM POWER PLANT ENGINEERING 

183. Cost of Engines. — In general the cost of engines per horse 
power diminishes as the size increases, but is of course governed by the 
style and workmanship. Average figures may be expressed as follows 
(Engr. U.S., Nov. 15, 1902, p. 750): 

Simple high-speed engines Cost in dollars = 300 + 8 X horse power 

Setting, high-speed engines Cost in dollars = 60 4- 0.75 X horse power 

Compound high-speed engines Cost in dollars = 1000 + 15 X horse power 

Simple low-speed engines Cost in dollars = 1000 + 10 X horse power 

Compound low-speed engines Cost in dollars = 2000 + 13 X horse power 

Setting, low-speed engines. Cost in dollars = 500 + 1.3 X horse power 

These equations were deduced from the curves in Figs. 163 to 165, 
which were plotted from the actual costs of a large number of engines. 
Rules for testing steam engines. — See Appendix C. 



CHAPTER X. 

STEAM TURBINES. 



184. Classification. — The following outline gives a classification of 
a few well-known steam turbines: 



Steam 

Turbines. 



Impulse 
Type. 



Reaction 
Type. 



Single J 
Stage. 1 



Multi- 
stage. 



Multi- 
Stage. 



De Laval. 




Sturtevant. 


Nozzle 




• Expansion. 


Curtis. 




Kerr. 




Terry. 




Rateau. 




Hamilton-Holzworth. 






Blade 


Parsons. 


• Expansion. 


Schulz. 




Allis-Chalmers. 





In the impulse type the steam is expanded by suitable means before 
doing useful work; that is, its potential energy is first converted into 
kinetic energy. In the reaction type the conversion is not complete, 
the expansion taking place partly before doing work upon the wheel 
and partly within the blades of the wheel itself. Thus the steam 
gives up a portion of its energy by direct impulse in impinging against 
the blades or buckets and the balance by reaction in leaving them. 
The impulse type may be either single- or multi-stage, depending upon 
the number of divisions in which expansion takes place, but the 
reaction type is always multi-stage. The single-stage impulse machine 
has one row of buckets or vanes mounted on the periphery of a 
revolving disk and one set of stationary nozzles. The peripheral velocity 
is very high, ranging in practice from 700 to 1400 feet per second. In 
the multi-stage impulse machines the expansion is divided between a 
number of stages, each one exhausting through suitably porportioned 
nozzles into the next succeeding stage. The steam velocity is thereby 
very much reduced, and the peripheral velocity may be considerably 
lower for good efficiency, ranging in practice from 200 to 600 feet per 
second. In the reaction type the steam flows through a large number 
of rows of blades alternately fixed and revolving. 

327 



328 STEAM POWER PLANT ENGINEERING 

184a. General Elementary Theory. — A given weight of steam at a 
given pressure and temperature occupies a certain known volume and 
contains a known amount of heat energy. If the steam is permitted to 
expand to a lower pressure without receiving additional heat or giving 
up heat to surrounding bodies it is capable of doing a certain amount 
of work which will be the same whether the expansion takes place in 
the cylinder of a reciprocating piston engine, a rotary piston engine 
or the nozzles and blades of a steam turbine. 

Let W = weight of steam, lbs. per sec. 

E = energy given up by 1 pound of steam, ft.-lbs. 
P x = initial pressure, lbs. per sq. in. abs. 
P n = final pressure, lbs. per sq. in. abs. 
H 1 = initial heat contents per lb., B.T.U. 
H n = final heat contents per lb., B.T.U. 

Then the heat available for doing useful work is 

If the steam expands against a resistance, as, for example, the piston 
of a reciprocating engine, the energy given up in forcing the piston for- 
ward may be expressed 

E x = 778 W (H t - H n ) ft.-lbs. (72) 

If the steam expands within a perfect nozzle the energy will be given 
up in imparting velocity to the steam itself, thus: 

E 2 =W^- ft.-lbs., (72a) 

2 2g 

in which V t = velocity of the jet in feet per second. 

If the velocity of the jet is retarded to V n feet per second, as by 
placing a series of vanes in its path, then the energy given up to the 

vanes (neglecting all losses) is 

V 2 — V 2 

E = W * n - (72b) 

2g 

If the jet is brought to rest by the vanes (neglecting all losses), then 
V n = and the energy given up is 

E 3 = W^. (72c) 



But E X =E 3 . Hence, 
from which 



778W(H l -H n )=W^-, 
& 9 



V x = 223.9 VX - H^* (73) 

* For most purposes it is sufficiently accurate to make 223.9 = 224. 



STEAM TURBINES 329 

The jet issuing from the nozzle is capable of exerting an impulse 

equal to F upon any object in its path, thus: 

WV 
F = i^- 1 lbs. (74) 

9 

If A = the area of cross section of the jet in sq. ft. and y = weight 
of steam, lbs. per cubic foot, then W = yAV 1} or 

F = *^Il lbs. (74a) 

9 

The reaction, R, of the jet against the nozzles is equal in value and 

opposite in direction to the impulse, or 

B = F = EX± = rlR. (74b) 

9 9 

The theoretical horse power developed by a jet of steam flowing at 
the rate of one pound per second may be expressed 

E v 2 —V 2 
H * P - = 550 = 2^X550' (75) 

in which y ^ = initial velocity of the j et> ft per sec> 

V n = final velocity of the jet, ft. per sec. 
Steam consumption per horse power hour: 

W-jg. (75a) 



Heat consumption, B.T.U. per horse power, per min. 

60 
in which ^ = k eat Q £ fa e iiq U i(j at p ressure p n . 



(75b) 



Impulse efficiency of the jet = equation (72b) ~ equation (72c). 
Thermal efficiency 



V 2 — V 2 

ft = v * (76) 



E t = ^ — ^2- See equation (68). (76a) 

«l — On 

Efficiency ratio or " kinetic " efficiency: 

2545 
F r = w \jj _ H , ■ See equation (71). (76b) 

Equations (72) to (76b) are general and are applicable to all turbines 
of whatever make. 

The more important types of turbines will be discussed separately 
and an application of above equations will be made in each specific case. 



330 



STEAM POWER PLANT ENGINEERING 




STEAM TURBINES 



331 



185. The De Laval Turbine. — Fig. 166 shows a section through a 
De Laval steam turbine and gear case and illustrates the principles of 
the single-stage " velocity " type. The turbine proper, to the right 
of the figure, consists of a high-carbon steel disk W fitted at the 
periphery with a single row of drop-forged steel blades and inclosed in 
a cast-steel casing. The disk is secured to a light flexible shaft and is of 
such a cross section that the radial and tangential stresses through- 
out its mass are of constant value. A flexible shaft is employed which 
allows the wheel to assume its proper center of rotation and thus to 
operate like a truly balanced rotating body.* The shaft is supported 
by three bearings, F, P, and 2V. N is self-aligning and carries the 
greater part of the weight of the disk. P is a flexible bearing, entirely 
free to oscillate with the shaft, and its only function is to seal the wheel 
casing against leakage. The power is transmitted through a steel 
helical pinion K mounted on the extension of the turbine shaft X, to 
two large gears E, E at a reduction in speed of about 10 to 1. The 
blades, Fig. 167, are made with a bulb shank and fitted in slots milled 




Fig. 167. De Laval Blades. 

in the rim of the wheel. The flanges, at the outer end of the blades, are 
brought in contact with each other and calked so as to form a continu- 
ous ring. The inlet and outlet angles of the blades are made alike and 
are 32 degrees for smaller sizes and 36 degrees for larger sizes. 

The operation is as follows: Steam enters the steam chest D, Figs. 
166 and 168, through the governor (shown in detail in Fig. 169) and 
is distributed to the various adjustable nozzles, varying in number 
from 1 to 15 according to the size of turbine. In the earlier types 
the nozzles were uniformly distributed around the circumference, but 
in the later types are arranged in groups. As illustrated in Fig. 168, 

* The shaft diameter for a 100-H.P. turbine is but 1 inch and for a 300-H.P. 
approximately 1{% inches. 



332 



STEAM POWER PLANT ENGINEERING 



the nozzles are placed at an angle of 20 degrees with the plane of 
the disk. The steam is expanded adiabatically in the nozzles to the 
existing back pressure before it impinges at high velocity against the 




De Laval Nozzle. 



blades. After giving up its energy the steam passes into chamber G, 
Fig. 166, and out through the exhaust opening. Fig. 169 gives the 
details of the governor and vacuum valves. Two weights B are 







Fig. 169. De Laval Governor. 






pivoted on knife edges A with hardened pins C bearing on the spring D. 
E is the governor body, fitted in the end of the gear wheel shaft K, 
and has seats milled for the knife edges A. The spring seat D is 



STEAM TURBINES 333 

held against pins A by spiral concentric springs, the tension on which 
is adjusted by a milled nut /. When the speed exceeds the normal, 
centrifugal force causes the weights to fly outward and overcome 
the resistance of the springs. This pushes pin G against bell crank L, 
which in turn closes the double-seated valve, thus throttling the 
supply of steam. To prevent racing in case the load is suddenly 
removed the vacuum valve T is added to the governor mechanism. Its 
operation is as follows: The governor pin G actuates the plunger H 
under normal conditions without moving the plunger relative to the 
bell crank. In case the load is suddenly removed, centrifugal force 
pushes pin G against bell crank L until it reaches its extreme position 
and the valve is nearly closed and little steam enters the turbine. If 
this does not check the speed, plunger G overcomes the resistance of 
spring M, and H moves relative to L, and its adjustable projection 
presses against valve stem T and allows air to rush into the turbine 
chamber through passage P. 

The power of the turbine depends upon the number of nozzles in 
action, and these can be opened or closed by a hand wheel on each. 
Each nozzle performs its function as perfectly when operating alone as 
when operating in conjunction with others. 

De Laval turbines are made in sizes ranging from 1J to 300 horse 
power, condensing and non-condensing, and are designed to regulate 
within an extreme variation of 2 per cent from no load to full load. 

The speeds vary from 10,600 r.p.m. for the largest size to 30,000 
r.p.m. for the smallest, the gearing reducing these to 900 and 3000 
r.p.m., respectively, at the shaft. 

The diameter of the wheel varies from 4 inches in the smallest tur- 
bine to 30 inches in the largest, thus giving peripheral velocities of 
from 520 to 1310 feet per second. 

De Laval Turbine : Prac. Engr., Jan. 1, 1910; Trans. A.S.M.E., 25-1056; Elec. 
World, July 29, 1905, p. 194, Oct. 26, 1901, p. 693; Eng. Rec, Oct. 19, 1901, p. 371; 
West. Elecn., June 4, 1904, p. 463; Machinery, Oct., 1904, p. 6, Nov., 1904, p. 123; 
Electrician, March 4, 1904; Power, Oct., 1905, p. 593. 

See Table 48 for results of tests of turbines of this type. 

186. Elementary Theory. — De Laval Turbine. — The maximum theo- 
retical power developed by a jet of steam flowing through a nozzle is 
dependent only upon the weight of steam flowing per unit of time and 
the initial velocity. Therefore the higher the initial velocity for a given 
rate of flow the greater will be the power developed and the higher the 
efficiency. 

* In Europe De Laval turbines are made as large as 3750 H.P. (See Power 
and Engr., May 10, 1910, p. 708.) 



334 



STEAM POWER PLANT ENGINEERING 



The maximum weight of steam discharged through a nozzle of any 
shape and for a given initial pressure is determined by the area of the 
narrowest cross section or throat. 

To obtain the maximum velocity at the exit or mouth, for a given rate 
of flow, the nozzle should be proportioned so that expansion to the 
external pressure into which the nozzle delivers shall take place within 
the nozzle itself. If expansion in the nozzle is incomplete, sound waves 
will be produced and there will be irregular action and loss of energy. 
On the other hand, if expansion in the nozzle is carried below that of 
the external pressure at the mouth, sound waves will be produced with 
subsequent loss of energy even greater than in the former case. 

Experimental and mathematical investigations indicate that the 
pressure at the narrowest section of an orifice or the throat of a nozzle 
through which steam is flowing falls to approximately 0.58 of the initial 
absolute pressure (with resultant velocity of about 1400 to 1500 feet per 
second) and any farther fall in pressure must take place beyond the 
narrowest section. Thus for back pressures greater than 0.58 of the 
initial (conveniently takes as f ), maximum exit velocity may be ob- 




^00?0MMmmm, 



Fig. 170. Theoretically Proportioned Expanding Nozzle. 



tained from orifices or nozzles of uniform cross section or with sides 
convergent. For back pressure less than 0.58 of the initial the nozzle 
must first converge from inlet to throat and then diverge from throat to 
mouth in order to obtain maximum velocity. Without the divergent 
portion of the nozzle the jet will begin to spread after passing the throat, 
and its energy will be given up in directions other than that of the 
original jet. 

Fig. 170 shows a section through a theoretically proportioned expand- 
ing nozzle. The cross section of the tube at any point n may be cal- 
culated by means of equation 

WS n 



A n 



V n 



(76c) 



STEAM TURBINES 335 

in which 

A n = area in square feet. 

W = maximum weight of steam discharged, lbs. per sec. 
S n = specific volume of the steam at pressure P n . 



For saturated steam S n = x n u 



n> 



in which x n = quality of steam at pressure P n after adiabatic expansion 
from pressure P v 
u n = specific volume of saturated steam at pressure P n . 

For superheated steam, see equation at bottom of page 131. 

V n = velocity of the jet, feet per second. 
V n may be determined from equation (73): 



V n = 223.9 VH l - H n . 

By substituting H n = heat contents corresponding to pressure 
P n = 0.58 P t in equation (73) and (76c) the area at the throat may be 
readily determined. The cross-sectional area for other points in the 
tube may be determined in a similar manner by assigning values of 
H n corresponding to the various pressures. 

In case of a perfect nozzle H x — H n represents the heat given up 
toward producing velocity by adiabatic expansion from pressure P x to 
P n . In the actual nozzle the frictional resistance of the tube serves to 
increase its dryness fraction, but in doing so it decreases the amount of 
energy the steam is capable of giving up towards increasing its own 
velocity. If y one-hundredths of the heat H t — H n is utilized in over- 
coming frictional resistance, then the resulting velocity will be 

V = 223.9 V(l -i^iH.-Hn). (76d) 

The quality of the steam after expanding to P n against the resistance 
will be higher by an amount 

t ,-, y (H ± — H n ) f 

l n = increase in quality = > (7oe) 

in which r n = heat of vaporization at pressure P n . 

The curves in Fig. 171, calculated by means of equations (76b) and 
(73), show the relationship between velocity, quality, pressure and 
kinetic energy for all points in a theoretically perfect nozzle expanding 
one pound of dry steam per second from an initial absolute pressure of 
190 pounds to a condenser pressure of one pound. 

The curves in Fig. 172 are based upon the experiments of Gutermuth 



336 



STEAM POWER PLANT ENGINEERING 



{Zeit. d. Ner. Ingr., Jan. 16, 1904) and show the effect of a few shapes 
of nozzles and orifices on the actual weight of steam discharged for 
various rates of initial and final pressures, the smallest section of the 
tube remaining constant. 

The nozzles of most commercial types of steam turbines are made 
with straight sides as in Fig. 168, so that only the area at the mouth need 



5000 



4500 



4000 



3500 



2000 



1500 



1000 



500 



2500 - 



THEORETICAL DESIGN OF A DIVERGENT NOZZLE 



190 

180 






























\ 




























170 
160 


\ 




























\ 
























S 


* 


\ 




























140 


\ 




























130 




\ 


























120 




\ 










4 


K^S 














110 




\ 








4* 




















\ 


























90 












Q 


J *L, 


> 










































70 
60 
50 
40 
30 






















































































J 


















* 


A 


























^> 










I 






























s 




























10 













5S 




















^5£, 







40,000 80,000 120,000 160,000 200,000 240,000 280,000 

kinetic energy of the jet, in foot pounds 
Fig. 171. 



100 
90 
80 
70 



be determined in addition to that at the throat in order to lay out the 
shape of the tube. 

Equations (73) and (76b) are general and are applicable to steam 
of any quality, wet, dry, or superheated. For steam initially dry and 
saturated Napier's rule offers a simple means of determining the area 
at the throat, thus: 



W = 



A . 



70 



for 



P n = or < - P x 



(76e) 



in which 



STEAM TURBINES 337 

W = 0.029 A VP n (P l -P n ) for P n > f P lt 

W = maximum weight of steam discharged, lbs. per sec. 

A = area at the throat, sq. in. 

P l = absolute initial pressure, lbs. per sq. in. 

P n = absolute back pressure, lbs. per sq. in. 



.06 



05 



.04 



03 



02 



.01 

















































































4 




















.00 
















V 






















\ 


\ 
















VJ 
























>3 


\ 


















"v 


.05 






































\ 


\ 








































\ 








.04 






._,.. 






r i 














-> 










P 

2 




\ 




■ " 


.03 












, 








i 






H 




3 










\ 




P i- 


■ " 


p r 


p- 

1 


^ 


P 






\ 




























* 


















, 
















.02 


zMtW, 






1 






2 












p— 

i 






1 




P 








r~ 


'■■. 




.01 




P = 132 Lb. Per Sq. In. Absolute 








l 




4 












A 


rea 


)fOi 


ifioc 


0.0 


555 S 


q.Ir 


. 

























.1 .2 .3 A .5 



.7 .8 .9 1.0 .1 .2 .3 A .5 .6 .7 .8 .9 1.0 



Fig. 172. Flow of Steam through Nozzles. 

Moyer ("The Steam Turbine," 1st Edition, p. 40) states that the ratio 
of the area of a correctly proportioned nozzle at the throat A to the area 
at any point A n is very nearly proportional to the ratio of the pres- 
sure at point A n to the initial pressure, or 

A P, 



Pn 



(76f) 



The entrance to the tube is rounded by any convenient curve. 

The length of the tube may be roughly approximated by the following 

formula: 

L = Vl5 A , (76g) 

in which 

L = length between the throat and mouth, in inches. 

A = area at the throat, sq. in. 

Practice shows that the cross section of a nozzle, whether circular, 
elliptical, square or rectangular (the latter with rounded comers), has 
very little influence on the efficiency provided the inner surfaces are 
smooth and the ratio of the area at the throat to that of the mouth is 



338 



STEAM POWER PLANT ENGINEERING 



cDrrectly proportioned. The velocity efficiency of a properly propor- 
tioned nozzle with straight sides is about 95 to 97 per cent, corresponding 
to an energy efficiency of 92 to 94 per cent, so that it is not considered 
worth while to attempt to follow the more difficult exact curves. 

Example : — Find the smallest cross section of a frictionless conically 
divergent nozzle for expanding one pound of steam per second from an 
absolute initial pressure of 190 pounds to an absolute back pressure 
of 2 pounds and find six intermediate cross sections where the pressures 
will be 70, 30, 14.7, 8, 4 and 2 lbs. respectively. Compare the velocity 
and energy of the jet issuing from this nozzle with those of an actual 
nozzle in which 10 per cent of the heat energy is lost in friction. 

From steam and entropy tables we find the values of H, x, u, for 
absolute pressures corresponding to 190, 0.58 X 190 = 110, 70, 30, etc., 
lbs. per square inch as follows (theoretical nozzle): 





H. 


X. 


u. 


5 = xu. 


P x = 190 


1197.3 


1.00 


2.405 


2.394 


P 2 = 110* 


1152.6 


0.960 


4.047 


3.878 


P,= 70 


1117.9 


0.932 


6.199 


5.775 


P 4 = 30 


1057.2 


0.887 


13.75 


12.27 


P 5 = 14.7 


1011.3 


0.857 


26.78 


22.95 


P." 8 


947.8 


0.834 


47.26 


39.29 


P 7 = 4 


935.6 


0.810 


90.4 


73.2 


P«= 2 


899.3 


0.788 


173.1 


137.0 



* P 2 = 0.58 Pi ( = pressure at throat). 



If entropy tables or charts are not available, values H l to H 8 and 
Xj to x s may be determined as outlined in equations (66b) to (67g). 

The different quantities for the theoretical nozzle will be calculated 
for the exit pressure P n = P s = 2 lbs. per sq. in absolute. 



V 8 = 223.9 VH, - H 8 

= 223.9 V1197.3 - 899.3 
= 3865 feet per second. 

E 8 = 778 (H t - H 8 ) 

= 778 (1197.3 - 899.3) 
= 232,000 foot-pounds. 

WS 
V 

1 X 137 



A* = 



3865 
= .0353 square foot. 



STEAM TURBINES 



339 



dg = y/pt^A = 13.58 VI 

= 13.56 V3353 
= 2.54 inches. 

WV a 



F» = 



9 
3865 



32.2 

= 120 pounds. 



THEORETICAL NOZZLE. 



Quantity < 


V 
Ft. per Sec. 


E 
Ft.-Lbs. 


A 

Sq. Ft. 


d 
Inches. 


F 
Pounds. 




(73) 


(72) 


(76c) 




(74) 




Pressures - 


110 
70 
30 
14.7 

8 

4 

2 


1,496 
1,995 
2,650 
3,053 
3,339 
3,624 
3,865 


34,767 
61,853 
107,485 
144,742 
173,207 
203,968 
232,000 


.00259 

.00269 

.00461 

.00745 

.0119 

.0202 

.0353 


0.693 

0.702 

0.919 

1.1 

1.46 

1.92 

2.54 


46.4 

61.98 

82.3 

94.8 

103.7 

112.5 

120.0 



In the actual nozzle these values will be modified because of the 
frictional losses. Thus for P n = 2 lbs., 



V s = 223.9 \/ (l - y) (H 1 - H 8 ) 

= 223.9 V(l - 0.1) (1197.3 - 899.3) 
= 3667 ft. per sec. 

E 8 = 778 (1 - 0.1) (1197.3 - 899.3) = 208,800 ft.-lbs. 

y (# i - h 8 ) 



^8 3*H ~l~lfi #8 1 



Qs 



= 0.788 + 



0.1 (1197.3 - 899.3) 



1021 



= 0.788 + 0.029 
= 0.817. 



A» = 



WxJuc 



_ 0.817 X 173.1 

3667 
= 0.0386 sq. ft., 



340 

from which 



STEAM POWER PLANT ENGINEERING 



d 8 = 2.66 inches. 

WV 8 3668 
8 a 32.2 



114 pounds. 



These various factors for all given pressures have been calculated in 
a similar manner and are as follows: 



ACTUAL NOZZLE. 



Quantities < 


V 
Ft. per Sec. 


E 
Ft. -Lbs. 


x'. 


A 

Sq. Ft. 


d 
Inches. 


F 

Ft. -Lbs. 




110 


1,420 


31,317 


.9658 


.00275 


0.711 


44.1 




70 


1,893 


55,632 


.9414 


.00286 


0.723 


58.8 




30 


2,515 


98,257 


.9026 


.00493 


0.951 


78.12 


Pressures ■ 


14.7 


2,894 


130,050 


.876 


.0080 


1.2 


98.8 




8 


3,168 


155,858 


.856 


.0127 


1.53 


98.4 




4 


3,438 


183,581 


.836 


.0220 


2.01 


106.8 




2 


3,667 


208,800 


.817 


.0386 


2.66 


114.0 



Many of these values may be determined directly from the Mollier 
or total heat-entropy diagram as described in Appendix H; in fact, the 
Mollier diagram has to all intents and purposes supplanted the steam 
tables in this connection. For superheated steam the diagram is 
extremely useful in avoiding laborious calculations. 




Fig. 172a. Velocity Diagram. Ideal Single-stage Impulse Turbine. 

Fig. 172a gives a diagrammatic arrangement of the blades in a 
De Laval turbine. The nozzle directs the steam against the blades with 
absolute velocity V x and at an angle a with the plane of the wheel XX. 
Since the wheel is moving at a velocity of u feet per second, the velocity 



STEAM TURBINES 



341 



v x of the steam relative to the wheel is the resultant of V t and u. The 
angle ( /? 1 between v x and XX will be the proper blade angle at entrance. 
If the blade curve makes this angle with the direction of motion of the 
wheel no shock will be experienced when the steam enters the blades. 
For convenience in construction the exit angle /? 2 is made the same as 
the entrance angle j3 v Neglecting frictional losses in the blade channels 
the relative exit velocity will be v 2 = v l} and the absolute velocity V 2 is 
the resultant of v 2 and u. The impulse exerted by the jet in striking 

W 

the vanes is — v v and its component in the direction of motion is 

W W 

— v t cos /?! = — (V l cos a — u). As the jet leaves the vanes the im- 

W W 

pulse is v 2 cos /? 2 .= (V 2 cos y + u). 

The total pressure acting on the vanes, or the actual driving impulse, is 
W 
9 



V, cos a 



(— V 2 cos y + u) 



W 



(V l cos a + V 2 cos y). 



(77) 



Equation (77) may also be expressed 

W 

p = _ . 2 (7, cos a -u). 
9 

The resultant axial force or end thrust is 

W 

F = — {V 1 sin a — V 2 sin y). 



(77a) 



(77b) 



Evidently if a = y and V\ = V 2 there will be no end thrust, since 
V l sin a — V 2 sin y will be zero. 
The work done is 

W 

Pu = — u (V\ cos a + F 2 cos y), (77c) 



or, using equation (77a) in place of (77), 

W 
Pu = — ' 2 u (V t cos a — u) 

W 

= — • 2 (uV, cos a— u 2 ). 

g I J 

By making the first derivative equal to zero 

d ( W ) 

j-) — 2 {uV 1 cos a — u 2 ) \ = V x cos a — 2 u = 0, 



or 



^ y t cos a. 



(77d) 



342 STEAM POWER PLANT ENGINEERING 

That is, for any nozzle angle a the work done, Pu, has its greatest 
value when u = \ V x cos a or y = 90°, whence 

Pu= W ^-cos 2 a. (77e) 

The work for any initial velocity V t becomes a maximum when a = 
and tt = J V v This condition can only occur for a complete reversal 
of jet and zero final velocity. Substitute a = and u = \ V x in 
equation (77d). 

TFT 2 
Pu = — — — » which is necessarily the same as equation (72c). 

In the actual turbine the various velocities will be less than those as 
obtained on account of the frictional resistance in the blades, and the 
velocity diagram should be modified accordingly. 

Example. Lay out the blades (theoretical and actual) for the nozzle 
in the preceding example, assuming that the jet impinges against the 
wheel at an angle of 20 degrees and that the peripheral velocity is 
1250 feet per second. 

Theoretical Case. 

Lay off V x = 3865 feet per second in direction and amount as shown 
in Fig. 172a and combine it with u = 1250 feet per second; this gives 
v v the relative entrance velocity as 2725 feet per second and /?, the 
entrance angle as 29 degrees. 

Lay off v 2 = v t at an angle /? 2 = & and combine with u; this gives 
V 2 , the absolute exit velocity, as 1740 feet per second. 

The theoretical energy available for doing work is 

W 

E = f g (V 1 >-V 2 >) 

= -7TT-A (3865 2 - 1740 2 ) = 185,000 ft.-lbs. 
64.4 

The difference between 232,000 and 185,000 = 47,000 ft.-lbs. is evi- 
dently the kinetic energy lost in the exhaust due to the exit velocity. 
The pressure exerted by the steam on the buckets is 

W 

P = — (V t cos a + V 2 cos 7) 

= ^ (3865 X 0.9397 + 1740 X 0.65166) 
= 148 pounds. 
The theoretical impulse efficiency is 



V t 2 - V 2 2 _ 3865 2 - 1740 2 _ 

V7 - ~ 3865 2 " U '' y/ 



STEAM TURBINES 



343 



The theoretical horse power per pound of steam flowing per second is 

185,000 



H.P. 



550 



= 336. 



Theoretical steam consumption per H.P.-hr. is 
3600 



336 



10.7 pounds. 



Actual Case. 

Proceed as in the theoretical case, using the actual absolute velocity 
V x = 3865 Vl - y = 3865 Vl -0.10 = 3667 feet per second in place 
of the theoretical value V t = 3870. Lay off V t = 3667 at an angle of 
20° as before and combine with u = 1250, Fig. 172b. 

U = 1250 




U = 1260 

Fig. 172b. Velocity Diagram as Modified by Friction Losses. 

The resultant v l = 2530 is the velocity of the jet relative to the wheel, 
and the entrance angle /? is found to be 29.7 degrees. The relative exit 
velocity v 2 will be less than v x because of the blade friction. 

Assume the loss of energy <f> from this cause to be 14 per cent; then, 
since the velocity varies as the square root of the energy, 

u, = v. Vl — d> 



= 2530 VI -0.14 
= 2346 ft. per second. 

The resulting absolute velocity V 2 is found from the diagram to be 
V 2 = 1405 ft. per second. 
Since the loss of energy in the nozzle is 

V 2_ {X _ y)V 2 



and that in the blade 



2g 

(1-0) TV 
2.9 



344 STEAM POWER PLANT ENGINEERING 

the remaining energy, deducting both losses in the nozzle and the 
blades, is 

= j^j (3865 2 - 0.1 X 3865 2 - 0.14 X 2530 2 - 1405 2 ) 

= 164,200. 
The losses due to windage, leakage past the buckets and mechanical 
friction must be deducted from these figures to give the actual energy 
available for doing useful work. Assuming a loss of 15% due to this 
cause, the work delivered is 

0.85 X 164,200 '= 139,570 ft.-lbs. 
The efficiency in the ideal case was found to be 0.797 and the available 
energy 185,000 ft.-lbs. 
The efficiency, deducting the loss due to friction, etc., is 
139,570 



185,000 
The horse power delivered is 

139,570 



0.797 = 0.60. 



= 254. 



550 
Steam consumption per horse-power hour is 

3600 tin A 

-T-— - = 14.2 pounds. 

254 r 

The heat consumption, B.T.U. per H.P., per minute is 
16.4 (1197.3 - 94) 

60 = 316 ' 

Assuming the revolutions per minute to be 10,000, the mean diameter 
of the wheel to give a peripheral velocity of 1250 ft. per second is 
1250 X 60 _ QO , + oc „. , 
1000 X 3.14 = 2 - 39 ft " ° r 28 ' 6 mcheS ' 
The determination of the height and width of vanes, clearance between 
nozzles and blades, etc., are beyond the scope of this work and the reader 
is referred to the accompanying bibliography. 

Blade Design for De Laval Turbines: Moyer, Steam Turbine, Chap. IV; Power, 
Mar. 17, 1908, p. 391. 

Flow of Steam through Nozzles: Jour. A.S.M.E. Mid. Nov., 1909, April, 1910, 
p. 537; Engineering, Feb. 2, 1906; Engr., Lond., Dec. 22, 1905; Eng. Rec, Oct. 26, 
1901; Power, May, 1905; Eng. News, Sept. 19, 1905, p. 204; French, Steam Tur- 
bines, Chap. XI; Pro. Inst. Civ. Engrs., Feb. 2, 1906. 

Design of Turbine Disks: Engr., Lond., Jan. 8, 1904, p. 34, May 13, 1904, p. 481; 
Jude, Theory of the Steam Turbine, Chap. XIII; Thomas, Steam Turbines, Chap. VI. 






STEAM TURBINES 



345 



Steam Turbine Efficiency: Power, Feb., 1906, p. 83; Jude, Theory of the Steam 
Turbine, Chap. VIII. 

Critical Velocity of Shafting: Jour. A.S.M.E., June, 1910, p. 1060; Power, Sept., 
1903, p. 484; Stodola, Steam Turbines, p. 177; Jude, Theory of the Steam Turbine, 
Chap. XVI; French, Steam Turbines, Chap. XV. 

Tests of De Laval Turbines: Eng. Rec., Aug. 2, 1902, p. 100; Am. Elecn., Aug., 
1905, p. 445; Engr. U. S., Aug. 1, 1905, p. 526; Eng. and Min. Jour., Nov. 3, 1904, 
p. 706; Machinery, Aug., 1904, p. 560; Eng. News, June 19, 1905, p. 62. 

187. Terry Turbine. — Fig. 173 shows a section through a Terry 
turbine, illustrating an application of the impulse type with two or 
more velocity stages. The rotor, a single wheel consisting of two steel 




Fig. 173. Section through Terry Steam Turbine. 

disks held together by bolts over a steel center, is fitted at its periphery 
with pressed steel buckets of semicircular cross section. The inner 
surface of the casing is fitted with a series of gun-metal reversing bucket 
arranged in groups, each group being supplied with a separate nozzle. 
The steam issuing from nozzle N, Fig. 174, strikes one of the buckets, 
B, on the wheel and, since the velocity of the buckets is comparatively 
low, is reversed in direction and directed into the first one of the revers- 
ing chambers. The chamber redirects the jet against the wheel, from 
which it is again deflected; this is repeated four or more times until the 
available energy has been absorbed by the rotor. Terry turbines are 
made in a number of sizes varying from 2 to 800 horse power, and operate 



346 



STEAM POWER PLANT ENGINEERING 



at speeds varying from 210 feet per second in the smaller machine to 
260 feet per second in the larger. These low speed limits compared 
with the speed of single-stage De Laval turbines are made possible by 
the application of the velocity stage principle in the use of the reversing 
buckets. The rotor of the smaller machine is 12 inches in diameter and 
runs at 4000 r.p.m., and that of the larger, 48 inches, running at 1250 




Fig. 174. Arrangement of Buckets and Reversing Chambers in a Terry Steam Turbine. 

r.p.m. Since the flow of steam into and from the buckets is in the plane 
of the wheel there is no end thrust. 

For a description of the Bliss, Dake, Sturtevant and Wilkinson steam 
turbines with results of tests see " Small Steam Turbines," by G. A. 
Orrok, Jour. A.S.M.E., May, 1909, and contributed discussion, Sept., 
1909. See also, " The Development of the Small Steam Turbine," 
Eng. Mag., Dec, 1908, and Jan., 1909. 

188. Kerr Turbine. — Fig. 175 shows a longitudinal section through, 
and Fig. 176 a sectional elevation of, a Kerr steam turbine. This 
turbine is of the impulse type and built on the principle of the 
Pelton water wheel, which it resembles in many respects. The rotor 
consists of a series of steel disks R, R, Fig. 175, mounted on a steel 
shaft. A series of drop-forged mild steel buckets of the double-cup 
type are secured to the periphery and riveted in dovetail slots. The 
stator is made up of a number of cast-iron diaphragms S, S, with cir- 
cular rims, which are tongued and grooved and when drawn together 
form a continuous cylinder. Square cold-rolled steel nozzle bodies 
N, N are expanded and beaded in the diaphragm near the rim and 
the nozzles screwed into them. The bearings B, B are of the oil-ring 
type; no thrust blocks are necessary, as each element is practically 
balanced. The operation is as follows: Steam enters the turbine at 
inlet A and passes through balanced throttling valve V (controlled by 
governor G) to the circular cored space H, H extending around the 



STEAM TURBINES 



347 







348 



STEAM POWER PLANT ENGINEERING 



entire casing. Space H, H acts as an equalizer and insures uniform 
admission to the first row of nozzles, where steam is partially expanded 
and the kinetic energy, imparted to the rotor through the medium of 
the buckets. The steam leaves the buckets at practically zero velocity 




Fig. 176. Kerr Steam Turbine: Sectional End Elevation. 



and is again expanded through the second set of nozzles. This process 
is repeated in each stage and the exhaust steam leaves the turbine 
at 0. Fig. 177 shows a diagrammatic arrangement of the governor. 
The governor weight is turned from solid steel and split into two piece, 7 





Ctr. Line of Shaft -? 



Section- A-A 

Fig. 177. Kerr Steam Turbine Governor. 



of semi-cylindrical form with the center of gravity near the center of the 
shaft. The weights are supported at three points. The hardened steel 
knife edge at B is of sufficient length for the stresses involved. The 
curve of rolling contact C is such that the bearing between the weigh 



STEAM TURBINES 



349 




TURBO-M-TERNATQR 




Fig. 178. Four-stage Vertical Curtis Turbo-Generator. Base Condenser T 



ype. 



350 



STEAM POWER PLANT ENGINEERING 



and the cam collar is always on the line of centers. The outward move- 
ment of the weights compresses the spring and operates, through lever 
connections, the balanced piston valve controlling the flow of steam. 
The movement of the center of gravity is indicated. 

The Kerr turbine is very simple in design, compact, noiseless, and 
low in cost of repairs. Its performance compares favorably with all 
other types of turbines of similar size and capacity. 

An 18-inch Kerr turbine direct connected to a multi-stage Worthing- 
ton centrifugal pump at the Armour Institute of Technology gives a 
steam consumption when running non-condensing comparable with 
that of high-grade non-condensing engines. 

Rateau Turbine : Trans. A.S.M.E., 25-782; Eng. Mag. T Oct., 1903, p. 49; St. Ry. 
Jour., April 18, 1903. 

Zolly Steam Turbine : Elec. Rev., Sept. 2, 1904. 

189. The Curtis Steam Turbine. — Figs. 178 to 183 show the general 
arrangement and a few details of the Curtis steam turbine, which is of 
the compound or multi-stage velocity type. The total expansion is 
carried out in one or more compartments or stages, each stage compris- 
ing a set of expanding nozzles and a wheel carrying two or more rows of 
buckets. A high initial velocity is given to the jet in each stage by 
expansion in the nozzles as in the De Laval, and the energy absorbed by 
successive action upon the series of moving and stationary vanes 
arranged somewhat as in the Parsons turbines, paragraph 192. In the 
latter, however, the difference in pressure between the two sides of each 
vane induces flow by continuous expansion, while in the former the 
moving vanes- in any one stage simply absorb the kinetic energy already 
created by expansion in the nozzle. The action is as follows : Steam enters 
stage (1), Fig. 180, through the first set of nozzles, and is partially 
expanded. With the resulting initial velocity it impinges against the 
first row of moving blades and gives up part of its energy, and is then 
deflected through the adjoining stationary blades to the next set of 
moving vanes, where its velocity is still further reduced, and so on until 
it has been brought practically to rest. From this stage the steam 
flows at reduced pressure through nozzles of stage (2), which are suffi- 
cient in number and in size to afford the greater area required by the 
increased volume. In expanding in these nozzles it acquires new 
velocity and gives up energy to the moving blades as before. This 
process is repeated through two to five stages, depending upon the size 
of turbine. Fig. 178 shows a partial section of a four-stage 5000-kilo- 
watt machine. R, R are sections through the revolving wheels, which 
in this particular turbine are nine feet in diameter and keyed to the 



STEAM TURBINES 



351 




352 



STEAM POWER PLANT ENGINEERING 



vertical shaft S. On the periphery of each wheel are bolted two rows 
of blades or vanes, with a stationary or intermediate row attached to 
the casing between them. The buckets are made of rolled nickel bronze, 
hammered to shape and finish. The roots are dovetailed into the holders 
and the tips are tenoned and riveted into a shroud ring, thus insuring 
positive spacing and a rigid construction. Between each pair of wheels 
is a stationary steam-tight diaphragm P, which contains the nozzles 
through which the steam is expanded from the preceding stage. It will 
be noticed that the buckets and nozzles increase rapidly in size in suc- 
ceeding stages as the pressure falls and the volume of steam increases. 







MOVING BLADE9 



STATIONARY BLADES 



MOVING BLADES 



STAGE 

No. 2 S 




NOZZLE DIAPHRAGM 



ffiHIffl 



!<<<<<<<<< 



) \i MOVING BLADES 

^^m STATIONARY BLADES 



n 



MOVING BLADES 



EXHAUST 

Fig. 180. 



The parts are so proportioned that the steam gives up approximately 

-of its energy in each stage, n representing the number of stages. 
n 

The number of stages and the number of vanes in a stage are governed 
by the degree of expansion, the peripheral velocity which is desirable 
or practicable, and by various conditions of mechanical expediency. 
The number of admission valves vary in number and in location with 
the size of turbine. The automatic stage valve G connects the first 
stage directly to a set of auxiliary second-stage nozzles. Thus the 
overload capacity is increased by widening the steam belt and not by 
admitting high-pressure steam into an intermediate stage as was for- 
merly the practice with Curtis turbines. This method of overload con- 
trol results in higher efficiency than with the older system. 



STEAM TURBINES 



353 



Curtis turbines appear to have a wider range of economical application 
than any other type, commercial sizes ranging from a small horizontal 
unit of 7 kilowatts rated output to vertical units of 20,000 kilowatts 
capacity on the continuous 24-hour basis. The smaller machines, 
1000 kilowatts and under, are usually of the horizontal type, and the 
larger units, 3500 kilowatts and larger, are of the vertical type. Be- 
tween 500 and 3500 kilowatts they are made both vertical and horizontal. 
All Curtis turbines are governed by " cutting-out nozzles"; that is, 
full initial pressure is maintained in all the nozzles that are open and 




Fig. 181. Section through Curtis Governor. 



the capacity of the machine is controlled by varying the number in 
operation. Units under 1500 kilowatts are ordinarily controlled by a 
mechanical valve gear and the larger units by an indirect or relay system. 
In the older types this relay system was electrically operated; in the 
modern machines the valves are hydraulically controlled. 

Fig. 181 shows a section through a typical Curtis governor. Speed 
regulation is accomplished by the balance maintained between the 
centrifugal force of moving weights A A and the static force exerted by 
spring D. The governor is provided with an auxiliary spring F, for 
varying its speed when synchronizing, the tension in which is varied by 



354 



STEAM POWER PLANT ENGINEERING 



a small pilot motor controlled from the switchboard. The movement 
of the governor weights is transmitted through rod C to arm H and by 
means of the latter to the controlling mechanism of the valve gear. 
Fig. 182 gives an assembly view of the mechanical valve gear as 




Fig. 182. Assembly of Mechanical Valve Gears for 300-Kw. Curtis Steam Turbine. 



applied to a 300-kilowatt unit. The valve stems extend upward through 
ordinary stuffing boxes and are attached to notched crossheads 8, 8. 
Each crosshead is actuated by a pair of reciprocating pawls or dogs, 
6, 6, the lower one of which closes the valve and the upper one opens it. 
The several pairs of pawls are hung on a common shaft which receives 
a rocking motion from a crank driven by the turbine shaft. The cross- 



STEAM TURBINES 



355 



heads have notches milled in the side in which the pawls engage to open 
or close the valve, the engagement being determined by shield plates 2, 
the positions of which are controlled by the governor through the medium 
of suitable levers. Shield plates 6 are set one a little ahead of the 
other to obtain successive opening or closing of the v^es. The 
pawls are held in position when not in contact with the shield plates by 
springs W. 

Fig. 183 gives a diagrammatic arrangement of the hydraulically 
controlled valve gear mechanism. The motion of governor g is trans- 
mitted through lever i to lever a of the pilot valve /. Pilot valve j 
controls the supply of oil (under pressure) in cylinder k the piston of 
which actuates rods I, I. The movement of rod I is transmitted through 



•Q 



p 





Turbine 



Fig. 183. 



Diagrammatic Arrangement of Hydraulically Operated Valve Gear, 
Curtis Turbine. 



rack m to a small pinion. This pinion is mounted on the end of a shaft 
fitted with a number of cams, one a little ahead of the other, each cam 
controlling the opening and closing of a steam valve through the medium 
of rocker arm/. As the load on the turbine increases the governor slows 
down and causes the cam shaft to rotate in a reverse direction indicated 
by the arrow points in Fig. 183. This causes a proportionate number 
of valves to be lifted and held open, the number increasing as the load 
increases, until all are open. Should the load continue to increase, as 
in the case of overload, the secondary valve opens as previously described, 
connecting the first stage with a set of auxiliary second stage nozzles. 
Only the nozzles in the first stage are controlled by the governor. 
Should the turbine run above normal speed the emergency stop valve 
automatically closes the admission of steam to the nozzles. This device 



356 



STEAM POWER PLANT ENGINEERING 



consists of a steel ring placed around the shaft between the turbine and 
the generator. This ring is eccentrically mounted and the unbalanced 
centrifugal force is balanced by a helical spring. When the predeter- 
mined speed is reached the centrifugal force overcomes the spring ten- 
sion and the ring moves in a still more eccentric position. In this posi- 
tion the ring strikes a bell crank lever which trips the throttle valve 









































































1 






































































■ 




































































5? 2 










































































































































M 3 










































































































































2* 




































































































































*5 































































































































































































12345G78 

Steam Belt Area 

Fig. 183a. Steam Belt Area in Five-Stage Curtis Turbine. 

and permits it to close by its own weight and the unbalanced pressure 
on the valve stem. 

In the Curtis turbine the area of the steam admission is limited to a 
small portion of the circumference in the first stages and does not 




OIL DRAIN 
OIL SUPPLY 



Fig. 183b. Step Bearing for Curtis Turbine. 

extend around the entire circumference until the last stage is reached. 
See Fig. 183a. 

The step bearing of a vertical machine is illustrated in Fig. 183b. The 
weight of the rotor is supported by oil under pressure forced between 
the bearing blocks M and P, thus permitting the shaft S to revolve 
on a film of oil. The smaller disk M is attached by dowels F to the 
main shaft. Carbon packing rings 0, are used above the bearing to 
prevent leakage, and adjustment is provided in the lower bearing block 



STEAM TURBINES 



357 



RELIEF VALVE, 



by means of set screws. The oil pressure varies from 150 to 750 pounds 
per square inch according to the size of machine, the higher pressures 
being used in the larger machines. 

Fig. 183c gives a diagrammatic outline of the oiling system. A tank, of 
sufficient capacity to contain all the oil and fitted with suitable straining 
devices and a cooling coil, is located at a level low enough to receive 
oil by gravity from all points lubricated. A pump draws oil from this 
tank and delivers it at a pressure about 25 per cent higher than that 
required to sustain the weight of the turbine in the step bearing. A 
spiral duct baffle connects the source of pressure to the step bearing and 
serves to regulate the oil supply to the 
lower end of the shaft. This source of 
pressure is also connected through a 
reducing valve to the upper oiling sys- 
tem of the machine, in which a pressure 
of about 60 lbs. to the square inch is 
maintained. This system, which includes 
a storage tank partly filled with com- 
pressed air, operates the hydraulic gov- 
ernor mechanism and supplies oil to the 
upper bearings. Delivery of oil to these 
bearings is regulated by adjustable 
baffles designed to offer resistance to 
the oil flow without forcing the oil to 
pass through any very small opening 
which might easily become clogged. A 
relief valve is provided to prevent the 
pressure in the upper part of the oiling 
system from rising above a desirable 
limit. Drain pipes from the upper bear- 
ings and from the hydraulic cylinder and relief valve all discharge into 
a common chamber, in which the streams are visible, so that the oil 
distribution can always be easily observed. At some point in the high- 
pressure system adjacent to the pump it is desirable to install a device 
to equalize the delivery of oil from the pump, as is done by the air cham- 
ber commonly used with pumps designed for low pressure. A small 
spring accumulator is furnished for this purpose, except in cases where 
weighted storage accumulators are used. In large stations where 
several machines are installed, a storage accumulator is desirable and 
can advantageously be so arranged that it will normally remain full, 
but will discharge if pressure fails, and in doing so will start auxiliary 
pumping apparatus. 




TO ACCUMULATOR 



Fig. 183c. Arrangement of Oiling 
System, for Curtis Turbine. 



358 



STEAM POWER PLANT ENGINEERING 





DIRECT 


CURRENT. 




Kw. 


R.p.m. 


Kw. 


R.p.m. 


15 
25 
75 


4,000 
3,600 
2,400 


150 

300 
500 


2,000 
1,800 
1,500 


ALTERNATING CURRENT. 


300 

500 

1,000 

1,500 


1,800 

1,800 

1,200 

900 


2,000 

3,000 

to 

20,000 


900 
I 600-750 



For the description of a typical steam turbine station equipped with 
Curtis turbines see Chapter XX. 

General Description of Curtis Turbines: Power, March, 1909; Engr. U. S., Jan. 1, 
1908, p. 115; Power & Engr., Feb. 25, 1908, p. 284, Feb. 25, 1908, March 3, 1908; 
Elec. Wld., June 17, 1905, p. 1136. 

Guide Bearings, Oil Distribution and Carbon Packing: Power & Engr., April 14, 
1908. 

Mechanical Valve Gear: Power & Engr., March 10, 1908, p. 356. 

Hydraulic Valve Gear: Power, March, 1909, p. 189. 

190. Elementary Theory, Curtis Turbine. — Fig. 184 gives a dia- 
grammatic arrangement of the blades and nozzles in the first stage 
of a two-stage Curtis turbine, each stage consisting of one set of nozzles 
and two moving and one stationary sets of blades. The object of 
employing a number of stages is to permit of a low peripheral velocity 
without reducing the efficiency. Since the velocity of steam varies as 
the square root of the kinetic energy, the theoretical stage velocity may 
be determined by dividing the maximum initial velocity by the square 
root of the number of stages, assuming that the entire velocity of the 
jet is abstracted in each stage. Thus if the turbine were constructed 
with four stages the theoretical stage velocity would be reduced from 
say 3600 feet per second to 1800 feet per second. In general, in order 
to reduce the stage velocity to V\ feet per second, the number of stages n 
may be determined from the equation 

V 2 
n = ~> (77) 

in which V = maximum initial velocity. 

Referring to Fig. 184: the steam is expanded in the first stage from 
pressure P l to P 2 and issues from the first set of nozzles with absolute 
velocity V lf striking the first set of moving blades at an angle a with 
the line of motion of the wheel. The resultant v x of V x and the 



STEAM TURBINES 



359 



peripheral velocity u, is the velocity of the steam relative to the vanes ; 
and the angle /? which the line v l makes with the line of motion of the 
wheel is the proper entrance angle of the blades for the first set. 
Neglecting friction the exit angle y will be the same as the entrance 
angle /?. The resultant of v 2 the exit velocity relative to the blade, and 
u, the peripheral velocity, is V 2 , the absolute exit velocity. 



Nozzles 




Fig. 184. Velocity Diagram, Curtis Turbine. 

Since the second set of blades is fixed and serves as a means of changing 
the direction of flow, the absolute velocity entering them is V 2 . The 
angle d formed by V 2 and the center line of the stationary blades is the 
proper entrance angle. Neglecting friction the absolute exit velocity 
will be 7 3 = V 2 , and the exit angle will be e = d. The steam flowing 
from the stationary blades strikes the second set of moving blades at 
an angle £ = d with absolute velocity V 3 . Combining V 3 with the 



360 STEAM POWER PLANT ENGINEERING 

peripheral velocity u we get v 3 , the velocity of the steam relative to the 
second set of moving blades. The angle 6, formed by v 3 and the line 
of motion of the wheel, is the proper entrance angle for the second set 
of moving blades. The resultant of v 4 (■= v 3 ) and u is V 4 , the absolute 
exit velocity for the first stage.* 

In the second stage the steam is expanded from pressure P 2 to that 
in the condenser and acquires initial velocity V a , leaving the last bucket 
with residual velocity V n - The theoretical velocities and blade angles 
for this stage may be found as above. 

Example: A two-stage Curtis turbine develops 800 horse power on a 
steam consumption of 12.5 pounds per horse-power hour, steam dry and 
saturated. Initial gauge pressure 135 pounds per square inch and back 
pressure 2 pounds per square inch absolute. Peripheral velocity 600 feet 
per second. The steam expands in the first stage from 150 pounds to 20 
pounds absolute, and in the second stage from 20 to 2 pounds absolute. 
Angle of the nozzles with the plane of rotation, 20 degrees. Compare the 
performance of the actual turbine with its theoretical possibilities. 
Actual turbine: 

Steam consumed per hour = 800 X 12.5 = 10,000 pounds. 

Steam consumed per second = 10,000 -5- 3,600 = 2.78 pounds. 

Horse power developed per pound of steam flowing per second, 
800 ■*- 2.78 = 288. 

Kinetic energy = 288 X 550 = 158,400 foot-pounds per second. 

Thermal efficiency, equation (68), 

2545 
E > " 12.5 (1191.1 - 94.4) = 187 Per Cent - 

Heat consumption, B.T.U. per horse power per minute, 

12.5 (1191.1 - 94.4) 

60 

T? 1 8 7 
Efficiency ratio = -=r = ^^ = 0.739. 

Ideal turbine: 

Stage velocities. 

The theoretical velocity in the first stage in expanding from a pres- 
sure of 150 pounds to 20 pounds absolute will be 

V l = 224\ / H l - H 2 

= 224 V1191.1 - 1042.9 

= 2727 feet per second. 

* In the actual turbine the velocities will be less than the theoretical on account 
of f rictional resistances in the nozzles and blades, and the velocity diagram must be 
modified as indicated in Fig. 172b and described on p. 



STEAM TURBINES 361 

In the second stage the steam expands from 20 pounds to 2 pounds 
absolute. 

7 = 224V#3-# 4 * 
- 224 V1042.9 - 914.8 
= 2530 feet per second. 

The kinetic energy per pound of steam in each division of the first 
stage will be: 

In the first set of moving blades, 

E = (V 2 — V 2 ) 

= ^ (2727 2 - 1620 2 ) = 74,722 foot-pounds per second. 

The value of V 2 is conveniently obtained from the velocity diagram. 
In the second set of moving blades, 



E=t^-7(V 2 2 -V 3 >) 



1 

64.4 

1 
64.4 



(1620 2 - 925 2 ) = 27,448. 



Total energy in first stage = 102,170. 

In a similar manner the total energy in the second stage will be found 
io be 93,700. 
Total for entire turbine = 195,870 foot-pounds per second. 
Theoretical horse power per pound of steam: 

Theoretical steam consumption per horse-power hour, 

3600 ini , 

— - = 10.1 pounds. 
Sob 

Heat consumption, B.T.U. per horse power per minute, 

10.1 (1191.1 - 94.4) 



60 
Thermal efficiency ratio, 



= 184. 



_ 1191.1 - 914.8 _ 
Er ~ 1191.1 - 94.4 " ° 253 - 

* In the actual turbine the heat contents H 2 , H 3 and H 4 will be greater than that 
of the ideal mechanism on account of frictional losses. See equation (76d). 



362 



STEAM POWER PLANT ENGINEERING 

SUMMARY. 



Horse power developed per pound of steam 
Steam consumption, pounds per H.P. hour 

B.T.U. consumed per H.P. per minute 

Thermal efficiency, per cent 

Efficiency ratio, per cent 



Actual Turbine. 



Perfect Turbine. 



288 


356 


12.5 


10.1 


288 


184 


18.7 


25.3 


73.9 





191. The Hamilton-Holzworth Steam Turbine. — Figs. 185 to 188 
give a general view and some of the details of the Hamilton-Holzworth 
turbine, which belongs to the compound multi-stage " velocity " type. 
The steam flows through the annular space between rotor and stator 



CONDENSER 

OR 

EXHAUST 

PRESSURE 




Fig. 185. Principles of Hamilton-Holzworth Steam Turbine. 

as in the Parsons, but differs from the latter in that expansion takes 
place only in the stationary vanes. The rotor consists of a number 
of steel disks of varying diameters riveted to both sides of steel hubs 
and fitted at the periphery with drop-forged vanes as shown in Fig. 
186. A tough steel ring is shrunk on the outside periphery of the vanes 
as indicated. The number of wheels and vanes is considerably less 
than in the Parsons type. The stationary vanes are fitted in steel 
disks as shown in Fig. 185, and the latter are located in grooves in 
the turbine casing. The vanes have a varying radial height increasing 
in the direction in which the steam flows. Sizes under 750 kilowatts 



STEAM TURBINES 



363 



have but one turbine casing, but larger sizes are divided into two, a 
high and a low-pressure turbine. The operation is as follows: Steam 
enters the high-pressure casing as indicated by arrows in Fig. 185, and 
passes through the first set of stationary vanes, extending around the 
whole periphery, which direct the steam at the proper angle against the 
wheel blades. In passing through the stationary vanes the steam is 
expanded down to the pressure in the first stage which is the same on 
both sides of the rotating disk. After giving up part of its energy, the 
steam expands again through the second set of stationary vanes to the 

pressure in the second stage, 
giving up energy to the second 
set of moving vanes. This pro- 
cess is repeated until the last 
stage is reached, from which 
the steam is discharged to the 
condenser in the simple turbine, 
or to the low-pressure steam 
chest in the compound turbine. 
In the low-pressure casing the 
steam is distributed in the same 
manner as in the high-pressure 
turbine. The diagram in the 
lower part of Fig. 185 shows 
the variation in steam pressure 
and velocity. The low-pressure 
front head is provided with an 
auxiliary nozzle which may be 
supplied with live steam in case 
of overload. The builders claim 
that since the pressures on both 
sides of the wheel are the same, 
no provision is necessary for axial balancing as in the Parsons standard 
turbine. 

Fig. 187 shows a sectional view of the bearing and stuffing box 
for the shaft at the point where it passes through the end of the turbine 
casing. The shaft is turned to a smaller diameter at its end and runs in 
a bushing G having a flange bearing against the inner side of the 
pillow block. At A is a cylindrical piece attached to and rotating with 
the shaft. This piece projects into an annular groove in the piece B, 
but it does not completely fill the groove and a circuitous passage 
is formed through which the steam must pass before reaching the 
stuffing box C. The object of the passage is to provide condensing 




DETAILS OP VANES. 
SECTION OF WHEEL. 

Fig. 186. Details of Vanes, Hamilton- 
Holz worth Steam Turbine. 



364 



STEAM POWER PLANT ENGINEERING 



surface so the steam itself will not reach the packing. The joint at the 
stuffing box is thus practically water-sealed. 



I 




Fig. 187. Details of Bearing, Hamilton-Holzworth Turbine. 

To prevent the oil from working into the turbine a bushing F is 
attached to -the shaft which throws off the oil into the space D by cen- 
trifugal force, where it drips down through a channel into a compart- 
ment in the pillow block. Any water 
escaping through the stuffing box is 
also collected in the same compart- 
ment. The bearing is oiled by a 
forced-oil system, the oil being sup- 
plied to the bottom of the bushing. 

Fig. 188 gives a diagrammatic view 
of the governor mechanism. A is the 
friction disk, L the roller, C the splined 
shaft with which the roller turns but 
upon which it is free to slide, bevel 
gears connecting shaft C with the 
throttle valve, and E a worm wheel 
driving disk A by means of a worm 
shaft F. At normal speed the gov- 
ernor sleeve is in mid position and 
roller L is at the center of the disk. 
If the turbine speeds up, however, the 
governor sleeve will rise, carrying with 
it the right-hand arm of lever T, which 
in turn will push the roller L a cor- 
responding distance downward. At 
the same time the cam H will be 
thrown to the right by contact with the roller R and by means of 
lever U will move the disk A and its shaft to the left, bringing it in 




Fig. 188. Governor Mechanism, 
Hamilton-Holzworth Turbine. 



STEAM TURBINES 365 

contact with roller L, thus imparting a rotary motion to the shaft C 
and closing the throttle valve until the turbine assumes normal speed, 
when the several parts assume the first position. If the speed is 
reduced below normal the operation is just the same except that the 
various motions are reversed. 

At an increase in speed of 2 J per cent above normal steam is cut off 
entirely. 

HamiUon-Holzworth Turbine: Am. Elecn., Oct., 1904, p. 549; Eng. Rec, Oct. 1, 
1904, p. 405; Power, Dec, 1907, p. 878, Nov., 1904, p. 659; Machinery, Nov., 1904, 
p. 134; Engr. U.S., Oct. 1, 1904, p. 690. 

192. Westinghouse-Parsons Steam Turbine. — Fig. 189 shows a section 
through a Westinghouse-Parsons multi-stage reaction turbine. In this 
type no nozzles are employed and expansion of the steam is effected 
by a series of stationary and movable blades. The rotor is a steel 
barrel or drum divided into three sections of varying diameter, upon 
the periphery of which bronze blades are radially inserted in dove- 
tailed grooves. The adoption of three sections of varying diameter 
has no bearing on the design of this machine but is merely for 
mechanical convenience. The blades increase in length and cross 
section from the high-pressure to the low-pressure end of each section. 
The stator is of cast iron and its inner surface is studded with rows of 
blades projecting radially inward and conforming in size with the adjoin- 
ing blades of the rotor. The relative positions of the blades in the rotor 
and stator are shown in Fig. 190. The operation of the turbine is as 
follows: Steam enters at S, Fig. 189, through poppet valve V, which is 
actuated by the governor shown in detail in Fig. 191, and flows through 
the annular space between rotor and stator to the exhaust opening 
at B. The entire expansion is carried out within this annular compart- 
ment and resembles in effect a simple divergent nozzle with the excep- 
tion that the dynamic relationship of jet and vane is such as to secure 
a comparatively low velocity from inlet to exhaust. The velocity 
varies from 150 feet per second at the high-pressure end to about 600 
feet per second as a maximum at the low-pressure end. The action 
of the steam on the blades is illustrated in Fig. 190. The steam strikes 
the first set of stationary blades as at P with initial velocity of about 
150 feet per second and is deflected against the moving blades immedi- 
ately adjoining. In passing from P to P t the steam is partly expanded 
and gives up a portion of its energy to the moving blades. The steam 
is deflected from P x to P n and thus has a reactive effect on the moving 
blades in addition to the impulse imparted at P v The total torque 
produced at the shaft in element A is therefore due to impulse from 



366 



STEAM POWER PLAXT ENGINEERING 







STEAM TURBINES 



367 



1 and reaction from 2. This process is repeated in each element of 
the turbine, the steam expanding as it flows from element to element 
in its passage to the condenser. The angular velocity of the rotor 



UJ 

-J 



5 CO 



Stationary 



ccuccuc 

HHHUD 



Blades 



"Moving Blades 



Vs 



Stationary 



ccccccccc 

)) )) )) B )) )) )) )) 



Blades 



Moving Blades 



Fig. 190. Flow of Steam in Parsons Turbine. 

varies from 3600 r.p.m. in a 400-kilowatt unit to 750 r.p.m. in the 7500- 
kilowatt size. Opposed to the three sets of blades the spindle also carries 
three rotating balance pistons P, P, Fig. 189, each of such diameter as 




Fig. 191. Governor Mechanism, Westingho use-Parsons Turbine. 

to exactly balance, through passage E, the axial thrust of the steam 
against its corresponding drum of blades. 

Steam enters the turbine intermittently as shown in Fig. 192, which 
represents indicator cards from a 1250-kilowatt turbine at various loads. 



368 



STEAM POWER PLANT ENGINEERING 



At light load the valve opens for a very short period and remains closed 
during the greater part of the interval. As the load increases, the period 
lengthens until finally, at about full load, the valve does not reach its 
seat at all, and continuous pressure is obtained in the high-pressure end 
of the turbine. 

The intermittent admission of steam is produced and controlled as 
follows: Lever T, Fig. 191, is given a reciprocating motion by an eccentric 
actuated by a worm and worm wheel on the main shaft. This motion 
is transmitted through lever H (with fixed fulcrum B) to lever A (with 
floating fulcrum D) and finally to pilot valve G. This reciprocating 
pilot valve admits puffs of steam from pipe to the under side of 
piston M, the rod R of which is attached to the admission valve V in 
Fig. 189. A spiral spring holds piston M in its lowest position until 




Fig. 192. Indicator Cards Showing Initial Pressure in a Westinghouse-Parsons 
* Steam Turbine. 



steam admitted by the pilot overcomes the spring tension and lifts the 
main valve from its seat, thereby permitting steam to enter the turbine. 
The fulcrum D of lever A is raised and lowered by the governor and 
therefore the pilot valve is controlled both by the motion of the eccen- 
tric and the motion of the governor. The eccentric keeps the pilot 
valve, and hence the main throttle, in constant oscillation, while the 
movement of the governor changes the limits of this motion. 

If an overload is sufficiently great to cause the governor balls to drop 
to their lowest position, the auxiliary or secondary valve V s , Fig. 189, 
begins to open and admits high-pressure steam to the later stage where 
the working steam areas are greater, thus increasing in proportion the 
total power of the turbine. The operation of this valve is the same as 
the main admission valve and is controlled by the governor. Fig. 193 
shows the details of this mechanism. The speed varies about 2 per 
cent from no load to full load. 



STEAM TURBINES 



369 



In the smaller size machines running above 1200 r.p.m., flexible bear- 
ings are employed to absorb the vibration incident to the critical velo- 
city. They consist of a nest of loosely fitting concentric bronze sleeves 
with sufficient clearance between them to insure the formation of a 
film of oil. In the larger machines running below 1200 r.p.m. a split 
self-aligning bearing is used instead of the flexible bearing. The ends 
of the casing are fitted with water-sealed glands of special design to 
prevent the escape of steam or inflow of air at the point of entry of the 
shaft. The water used for sealing them is small in quantity and may 
be returned to the feed-water system. . 




Fig. 193. By-Pass Valve, Westinghouse-P arsons Turbine. 



Double-flow Type. — In reaction turbines of the single-flow type, 
as illustrated in Fig. 189, the high-pressure portion dealing with the 
high-pressure incoming steam is the least efficient. This is due to 
the fact that the blade lengths are approximately proportional to the 
specific volume of the steam, and consequently the initial expansion 
in the turbine requires blade passages of very small dimensions. This 
results in greater leakage past the tips of the blades than in the low- 
pressure elements where the blades are long. Again, in the single-flow 
type the high-pressure balance piston occupies fully one-half of the 
total balance piston length of the shaft, while the low-pressure piston 
is 2\ times the high-pressure diameter, so that balance pistons occupy 
a large portion of the total bulk of the machine. By making the high- 
pressure element of the impulse type and by arranging the low-pressure 
reaction elements on either side as illustrated in Fig. 193b the efficiency 
may be increased and the bulk of the turbine may be greatly decreased. 



370 



STEAM POWER PLANT ENGINEERING 



There are two rows of moving buckets upon the impulse wheel with an 
intermediate set of reversing blades, the operation being practically 
the same as in the first stage of a Curtis turbine. The drop in pressure 
in the nozzles is such that approximately 20 per cent of the total energy 
developed is absorbed by this impulse element. After leaving the im- 
pulse element the steam divides, one portion passing directly to the 
low-pressure blading at the left, while the rest passes through the hollow 
shell of the rotor to the similar pressure blades upon the right. As 
these sections are equal and symmetrical they counterbalance each 
other, and the balance or " dummy " pistons may be dispensed with. 
The advantages of the double-flow type over a single-flow unit of equal 
capacity are: (1) reduction of nearly 50 per cent in the shaft span 



willl 





Plan View 



Side View 



End View 



Fig. 193a. Method of Fastening Blades in Westinghouse-Parsons Turbines. 

between bearings; (2) the diameters of the casing and rotating part 
are more uniform, thus tending to greater rigidity; (3) a reduction of 
about 70 per cent in the bulk of the main parts of the machine, and (4) 
internal stresses due to high-pressure and high-temperature steam are 
avoided by isolating the incoming steam, without separate nozzle 
chambers. Westinghouse-Parsons turbines are made in a number of 
sizes, varying from 400 kw. to 15,000 kw. In Europe, however, 
Parsons turbines are made as small as 20 kw. and as large as 25,000 
H.P. In sizes up to 3500 kw. the single-flow turbine has established 
itself as the most suitable prime mover, but for larger sizes the double- 
flow is given preference. The double-flow turbine is admirably adapted 
to low-pressure work. Fig. 194b shows a section through a Westing- 
house-Parsons double-flow low-pressure turbine. For results of tests 
of Parsons and W T estinghouse-Parsons turbines see Table 48. 

Double-flow Turbine: Power & Engr., March 16, 1909, Aug., 1908, p. 471 ; Eng. 
Rec., May 30, 1908, p. 693; Elec. Review, June 26, 1908, p. 1089. 

10,000-kw. Westinghouse-Parsons Double-flow Turbo Generator for the Metropolitan 
Street Railway Company, Kansas City, Kansas: Power & Engr., May, 17, 1910, p. 890. 



STEAM TURBINES 



371 




$ 

03 



372 



STEAM POWER PLANT ENGINEERING 



192a. Allis-Chalmers Steam Turbine. — Fig. 193c shows a section 
through an Allis-Chalmers standard steam turbine, which is of the 

Parsons type but differs 
from the original Parsons 
machine and the Westing- 
house-Parsons construc- 
tion principally in manu- 
facturing details. In the 
older Parsons type, three 
balancepistons are placed 
at the high-pressure end. 
In the Allis-Chalmers de- 
sign, the larger piston is 
placed at the low-pressure 
end of the rotor, behind 
the last row of blades, 
the other two remaining 
at the high-pressure end. 
This construction per- 
mits of a smaller balance 
piston and allows a smaller 
working clearance in the 
high-pressure and inter- 
mediate cylinders. In the 
Allis-Chalmers turbine 
the roots of the blades are 
dovetailed and fitted into 
a foundation ring, and the 
tips are encased in a chan- 
nel-shaped shroud ring,, 
thereby insuring a rigid 
and positively spaced con- 
struction. The governor 
is of the Parsons type, ex- 
cept that the main valve 
and pilot valve are actu- 
ated by hydraulic instead 
of steam pressure. The 
bearings are of the self- 
adjusting ball and socket pattern and are kept " floating in oil " by 
a small pump geared to the turbine shaft. The oil is passed through 
a tubular cooler with water circulation after it leaves the bearings and 
is used over and over again. 




STEAM TURBINES 



373 



193. Elementary Theory, Parsons Turbine. — Fig. 194 gives a dia- 
grammatic arrangement of fixed and stationary blades in the first 
stages of a multi-stage ideal reaction turbine. The steam enters the 
stationary blades at practically zero velocity and is there partially 
expanded and impinges against the movable blades at velocity V v 
part of the energy of the steam being thus absorbed. In passing 
through the movable blades the steam is still further expanded and 
leaves at an absolute velocity V 2 , exerting an additional pressure on 
the blades from the reaction. The steam enters the second set of 
stationary blades with velocity V 2 and is still further expanded to 
velocity V 3 , and so on. 

"P=150 




W=300 



Fig. 194. Velocity Diagram. Westingho use-Parsons Turbine. 

The energy imparted to the steam in the first set of stationary blades 



is 



E,= 



W 



(78; 



V t = absolute velocity of the steam leaving the blades. 
The energy imparted to the steam in the first set of moving blades is 

W 



E n 



2g 



(v 2 2 - v 2 ). 



(79) 



i\ = relative velocity of the steam entering the moving blades. 

v 2 = relative velocity of the steam leaving the moving blades. 

The total energy acquired by the steam in the first stage is 

E, + E 2 . 

The energy converted into work in this stage is 

WV 2 
E = E t + E 2 - V^ 
2 g 

V 2 = absolute velocity of the steam leaving the moving blades. 
Each stage may be analyzed in a similar manner. 



(80) 
(81) 



374 STEAM POWER PLANT ENGINEERING 

Example: A Westinghouse-Parsons turbine develops 1000 horse 
power on a steam consumption of 12 pounds of steam per horse-power 
hour. Initial steam pressure 150 pounds per square inch absolute; 
back pressure 1 pound per square inch absolute; drop in pressure in 
each set of fixed and moving blades 15 pounds per square inch; 
peripheral velocity 300 feet per second; a x = a 2 = 30 degrees. Com- 
pare the performance of the actual and ideal turbine. 

Actual turbine: 

Steam consumed per hour, 

1000 X 12 = 12,000 pounds. 

Steam consumed per second, 

12,000 ^ 3,600 = 3.33 pounds. 

Horse power developed per pound of steam flowing per second, 

1000 -^ 3.33 = 300. 

Kinetic energy per pound of steam, 

300 X 550 = 165,000 foot-pounds per second. 

Thermal efficiency, 

g '- 12(llSS-70) - 18 - <>P " 0mt - 

Heat consumption, B.T.U. per horse-power hour per minute, 

12(1191.2-70) _ 221 
60 
Efficiency ratio, 

Ml = 1^1 = 67.5 per cent. 

E r 28 F 

Ideal turbine : 

The velocity imparted to the steam in the first set of stationary 
blades due to the drop from 150 to 135 pounds per square inch is 



V t =224 VH 1 - H 2 



=224 V1191.2- 1182.4 

= 662 feet per second. 

Lay off the value of V r in direction and amount and combine with u, 
the peripheral velocity, Fig. 194. The resultant is v lf the velocity 
of the steam relative to the blades. The angle between v x and the line 
of motion of the wheel will be the angle with the blade at entrance. 

From the velocity diagram, 

v* = 429. 



STEAM TURBINES 375 

E 2 , the energy given up by one pound of steam in expanding from 
135 to 120 pounds, is 

E 2 =77S (H 2 -H 3 ) 

= 778 (1182.4-1172.8) 

= 7468 foot-pounds per second. 

Substitute v x = 429 and E 2 = 7468 in equation (79), 

7468 = — (v 2 2 - 429 2 > 
64.4 V 2 ; 

v 2 = 816 feet per second. 

The resultant of v 2 and u is V 2 , the absolute velocity of the steam 
leaving the moving blades of the first stage. From the diagram, 

V 2 = 573 feet per second. 

The energy converted into work in the first stage is determined by 
substituting the proper values in equation (81), thus: 

E = (662 2 + 816 2 - 429 2 -573*) -i- 

64.4 

= 9200 foot-pounds per second. 

The various stages may be analyzed in a similar manner. 

The theoretical output of the entire turbine per pound of steam will 
be that corresponding to adiabatic expansion from a pressure of 150 to 
1 pound absolute. 

E= 77S(H t - H n ) 
= 778(1191.2-877) 
= 244,447 foot-pounds per second. 

Horse power per pound of steam, 

H.P. = ?M£17 _ 445 
550 

Steam consumption per horse-power hour, 

3600 -, , 

=8.1 pounds. 

445 

Thermal efficiency, 

E = 1191.2-877 

r 1191.2-70 

= 28 per cent. 



376 



STEAM POWER PLANT ENGINEERING 



194. Low-pressure and Mixed-pressure Turbines. — A promising 
field for the steam turbine is in its application as a secondary or low- 
pressure unit in connection with non-condensing or condensing engines, 
or, combined with a regenerator, in connection with engines using steam 
intermittently. Numerous examples may be cited showing great gains 
in both capacity and economy in existing power plants involving the 




Fig. 194a. Low-pressure Turbine Installation at the 59th Street Station of the 
Interborough Rapid Transit Company, New York. 



abandonment of but a negligible part of the equipment and accom- 
plishing this result with a minimum additional investment. The most 
notable installation (June, 1910) of low-pressure turbines to con- 
densing reciprocating engines is at the 59th Street Station of the 
Interborough Rapid Transit Co., New York. Three of the nine 7500-kw. 
Manhattan-type compound Corliss engines have been equipped with 
Curtis three-stage, low-pressure turbo-generators of equal capacity, and 
provision is made for the installation of six additional units. The low- 
pressure turbine is installed between the exhaust of the low-pressure 
cylinders and the condenser as shown in Fig. 194a. Running with the 



STEAM TURBINES 



37T 




i^^^^^^^^^i 



y////////////rf& 




378 



STEAM POWER PLANT ENGINEERING 



engine the low-pressure turbine generator carries a variable load without 
governor regulation. The turbine generator takes care of the speed 
by automatically taking such a load as will keep the frequency in unison 




Engine Load K.W. 
3000 | 4000 | 5000 



Fig. 194c. Performance of 7500-Kw. Engine at 59th Street Station of Interborougli 
Rapid Transit Company, New York, with Varying Receiver Pressure. 



19 



18 



K 17 

u 
8.15 



13 



a « 

10 

















































Act 


lal 1 


'est) 


En§ 


;ine 








































































_IU_ 


Jarai 


itee) 


Hif 


f h Pressu 


re r J 


■"urbine 










































































i> 














_^A 






■ 














fuit 


iue 


















I*" 


Co 


nbic 


ed 1 


:ngu 


ie~ai 


id-t 


jw-tr« 3 = 






































A:- Constant Nozzle Pressure 

@ 16 #■ abs. 
B:- Variable Nozzle Pressure 
All Nozzles Open 
























































Cor 

Tur 


rectec 
sine £ 


I for 
team 


Moist 


ure li 


i 



7000 8000 



9000 



10000 11000 12000 13000 
Unit Load JC.W. 



14000 15000 16000 



Fig. 194d. Comparison of Economy Curves: 7500-Kw. High-pressure Turbine, 7500-Kw. 
Engine and Combined Engine and Low-pressure Turbine at the 59th Street Station 
of the Interborougli Rapid Transit Company, New York. 

with that of the engine-driven generator. The turbine is equipped 
with the usual emergency speed limit attachment for cutting off the 
steam supply should the speed exceed a predetermined limit. The 



STEAM TURBINES 



379 



performance of one set of engines, a high-pressure turbine of the equiva- 
lent total capacity, and that of the combined engine and low-pressure 
turbine, are illustrated in Fig. 194d. The conclusions drawn from an 
exhaustive series of tests at this station are that the addition of low- 
pressure turbines effected : 

a. An increase of 100 per cent in maximum capacity of plant. 

b. An increase of 146 per cent in economic capacity of plant. 

c. A saving of approximately 85 per cent of the condensed steam for 
return to the boiler. 

d. An average improvement in economy of 13 per cent over the 
best high-pressure turbine results. 

e. An average improvement in economy of 25 per cent (between the 




vfc:zGzzzz 



Fig. 195. Rateau Low-Pressure Steam Turbine Installation. 



limits of 7000 kw. and 15,000 kw.) over the results obtained by the 
engine units alone. 

/. An average unit thermal efficiency of 20.6 per cent between the 
limits of 6500 kw. and 15,500 kw. 

Low-pressure turbines are frequently installed in connection with 
regenerative accumulators, to rolling-mill engines, steam hammers, and 
other appliances using steam intermittently, and have proved to be 
paying investments. A typical installation of this character is to be 
found at the South Chicago Division of the International Harvester 
Company. The front elevation of the turbine and regenerator installa- 
tion is shown in Fig. 195 and the general arrangement of the regenerator 
is shown in Fig. 196. The regenerative accumulator is intended to 
regulate the intermittent flow of steam before it passes to the turbine. 



380 



STEAM POWER PLANT ENGINEERING 



The steam collects and is condensed as it enters the apparatus and is 
again vaporized during the time when the exhaust of the engines dimin- 
ishes or ceases. 

The regenerator consists of a cylindrical boiler-steel shell divided 
into two similar chambers by a central horizontal diaphragm. In each 
compartment are a number of elliptical tubes A, each of which is per- 
forated with a number of f-inch holes. The spaces surrounding the 
tubes and, under certain conditions, the tubes themselves are filled 
with water to a height of about four inches above the top of the upper 
tubes. Baffle plate B serves to separate the entrained moisture from 
the steam. The operation is as follows: Exhaust steam enters the 
apparatus at N, passes to the interior of the elliptical tubes, and escapes 




Fig. 196. Rateau Regenerator Accumulator. 

into the steam space through the perforations and thence to the turbine. 
When the supply of steam from the main engine ceases, the pressure in 
the regenerator decreases, the water liberates part of the heat it has 
absorbed and a uniform flow of low-pressure steam is given off. The 
continued demand of the turbine reduces the pressure in the accumula- 
tor and causes the steam still retained in the tubes to escape, thereby 
maintaining the circulation of the water (indicated by arrowheads) and 
facilitating the liberation of steam. Suitable valves regulate the limits 
of pressure in the accumulator and prevent the return of water to the 
main engine. Low-pressure turbines develop one electrical horse- 
power hour on a steam consumption of about 30 pounds with initial 
pressure of 15 pounds absolute and a back pressure of 1.5 pounds 
absolute. Fig. 197 gives the performance of the 500-kilowatt Rateau 
turbine at the International Harvester Works, South Chicago, I., 



, 



STEAM TURBINES 



381 



average initial pressure of 16 pounds absolute, condenser pressure 
1.5 pounds absolute. 

Low-pressure turbines equipped with special expanding nozzles, or 
the equivalent, to receive steam at high pressure direct from the boilers 
are known as mixed pressure turbines. With this construction the 
full power of the turbine can be developed with (1) all low-pressure 
steam, (2) all high-pressure steam, (3) any proportion of high and low 



Reinforced Cinder Concrete 




Fig. 196a. Typical Double-deck Installation, Fort Wayne and Wabash Valley 
Traction Company, Spy Run Station. 



pressure steam. In the Curtis mixed-pressure turbine this transition 
from all low pressure to all high pressure, through all the conditions 
intermediate between these extremes, is provided for automatically 
by the turbine governor; a deficiency of low-pressure steam causes the 
high-pressure nozzles to open automatically. With this arrangement 
it is not necessary for purposes of economy to proportion exactly the 
low-pressure turbine to the amount of exhaust steam available, but 
within limits it may be made as large as the load demands. 

Low-pressure Turbines: Power & Eagr., July 6, 1909, p. 1, Nov 30, 1909, p. 905; 
Prac Engr. U. S., Mar. 1, 1909, p. 169; Eng. Mag., Apr. and May, 1907; Iron Age, 
Jan. 7, 19*09. 



382 



STEAM POWER PLANT ENGINEERING 



195. Advantages of the Steam Turbine. — The principal advantages 
of the steam turbine are (1) simplicity; (2) economy of space and 
foundation; (3) absence of oil in condensed steam; (4) freedom from 
vibration; (5) uniform angular velocity; (6) large overload capacity; 
and (7) high efficiencies for large variations in load. The reciprocating 
engine is well adapted for pumping stations, direct-current generators, 
compressor plants, hoisting engines, and the like, requiring low angular 
velocity, but its place is being rapidly taken by the steam turbine for 
alternating-current dynamos, centrifugal pumps and blowers, requiring 
high angular velocity. 



— i rt mi i mi hi m 1 1 1 1 1 1 1 1 1 1 1 1 1 i — r 


i ii 1 1 ii 1 1 1 1 M ii i ii ii 1 1 ii i ii 1 1 1 1 














li ' \ 1 ! 1 


1 ^ 1 


1 \ t 


1 ^ 


di x _:::_ v v ±::::::_ : ::: 


^ i \ ' - 


r; n _ it- \_ X_ _ _ J 


50 - - _ s __ _ _ _ j---- ---- 


v t ::::: ::: 


\ _: :: 


\ it _:: 




_ 4^_ _ s _ _ _ __ 


\ : : _::: 


5v * -- 


'ST ~Z 


if) t "v ./ 




. % „? : z: 




: _ **« ,s" : ~: : 


*>* »<i 






























— LU L ±^± Z-.JLZ 



3U0 400 500 600 700 800 900 

H.P. at Switchboard 

Fig. 197. 

196. Simplicity. — Although composed of a large number of parts 
as compared with a reciprocating engine of the same capacity, there are 
few moving parts and rubbing surfaces. The only contact between 
rotor and stator is in the main bearings, and the problem of lubrication 
is therefore a simple one. The absence of pistons, stuffing boxes, dash 
pots, etc., reduces the cost of maintenance and attendance to a minimum 
and limits the possibility of leakage. 

197. Economy of Space and Foundation. — Fig. 198 (J. R. Bibbins, 
Power, January, 1905) shows the relative floor space required for different 
types of slow-speed reciprocating engines and Westinghouse-Parsons 
steam turbines. The floor space required by the Curtis turbine is con- 
siderably less than for the Parsons type, as it is vertical with the gener- 
ator mounted above the turbine. Vertical three-cylinder compound 



STEAM TURBINES 



383 



Corliss engines of the New York Edison type require the least floor 
space of any large slow-speed reciprocating engines, but take up about 
twice the space of a Parsons turbine installation of the same size. 
With high-speed engines of the Willans central-valve type the com- 
parative economy in space is less marked. The curves refer to the 
space occupied by engine and generator alone, whereas in a modern 
turbine installation with surface condenser the condenser equipment 
may occupy as much space as, if not more than, the turbine proper, 
and considering the small space taken up by the barometric condenser, 
such as is used in connection with the New York Edison engines, the 
economy is not so evident. In recent turbine installations the baro- 











































1 




1 










































































































































COMPARATIVE FLOOR SPACE FOR 












































REPRESENTATIVE TYPES OF PRIME MOVERS 


















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Based on Overall Dimensions 
of Generating Unit 





































































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Fig. 198. 



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metric condenser is finding much favor, and in such instances the curves 
may be taken to indicate the relative floor space for the entire equip- 
ment of prime mover and auxiliaries. 

The weight of the steam turbine is very small compared with a recip- 
rocating engine of the same horse power. The New York Edison engines 
and generators weigh more than eight times as much as a turbine 
installation of equal capacity. The turbine for this reason, and also 
because of the total absence of vibration, requires a relatively light 
foundation. In many instances the foundation consists of steel 
beams with concrete arches sprung between them resting upon the 
floor, and the basement underneath may be used for the condenser 
instead of the massive foundation required for the reciprocating engine. 



384 STEAM POWER PLANT ENGINEERING 

198. Absence of Oil in Condensed Steam. — As the steam turbine 
requires no internal lubrication, oil does not come in contact with the 
steam, and the condensed steam from the surface condensers is available 
for boiler-feeding purposes without purification. In many cases the 
re-use of condensed steam effects a large saving in cost of feed water 
and in expense for maintenance and cleaning of boilers. The amount 
of entrained air is reduced to a minimum and consequently the work 
of the air pumps lessened. 

199. Regulation. — The variable pressure at the crank pin of a 
reciprocating engine necessitates the use of a heavy fly wheel to keep 
the instantaneous angular fluctuation within practical limits. In the 
steam turbine the motion is purely rotary and a fly wheel is not neces- 
sary. In the former there are always instantaneous variations in 
velocity during each revolution, even with constant load, while in 
the latter the speed is practically constant. A number of published 
tests of Parsons and Curtis turbines show an average fluctuation of 
2 per cent from no load to full load and 3 per cent from no load 
to 100 per cent overload. Although closer regulation than this is 
possible, it is not deemed necessary, particularly in alternating- 
current work where a comparatively wide range is desirable for 
parallel operation. 

200. Overload Capacity. — A particular advantage of the turbine 
over the reciprocating engine lies in its greater overload. capacity and 
higher economy at overloads. The maximum economy of the average 
reciprocating engine lies between 0.75 and full load, whereas the 
turbine reaches its maximum at about 25 per cent overload. Thus 
a single turbine unit may economically take the place of two or 
more reciprocating units for a variable load. A turbine may be 
readily operated at 100 per cent overload, while the ordinary engine 
reaches its maximum capacity at about 50 per cent overload. In 
central lighting and power stations where there are one or more sharp 
peak loads of short duration, this extreme overload capacity is of 
marked importance. 

201. Efficiency and Economy. — As far as steam consumption is 
concerned there is practically no difference between the performance 
of a high-grade piston engine and that of a first-class turbine for sizes 
under 2000 kw., the choice depending more upon rotative speed, over- 
load capacity and space requirements than upon the heat economy. 
For sizes over 2000 kw. the fuel consumption lies in favor of the 
turbine. A comparison of Fig. 148, showing typical economy curves of 
high-speed single- valve, non-condensing engines, and of Fig. 198c, show- 
ing similar curves for small non-condensing turbines, is somewhat 



STEAM TURBINES 



385 



in favor of the piston engine, though the difference is small; whereas a 
comparison of the turbine and engine curves in Fig. 194d, showing the 
performance of very large units, is decidedly in favor of the turbine. 



W 3000 






























































































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Steam Pressure 175* 
Back Pressure .68* 
Superheat 60° 
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30 



40 



50 00 70 

Brake Horse Power 



80 



90 



Fig. 198a. Typical Performance of a 90 H.P. Terry Steam Turbine. 

Any number of individual tests may be cited showing superiority in 
fuel consumption of the piston engine over that of a turbine cf equiva- 
lent capacity and vice versa, but when the machines are designed for 



IP.14 
12 



5000 



G000 



rooo 



8000 9000 
Load in K.W. 



10000 



11000 



12000 



Fig. 198b. Typical Performance of 9000 Kw. Curtis Turbine; 200 Lbs. Gauge 
Pressure, 125° Superheat, 29 Inches Vacuum. 



the same operating conditions the results are practically the same for 
all sizes under 2000 kw. Tables 39 to 43 give the general condition 
of operation and the steam consumption of exceptionally good piston 
engines of various sizes and types, and Table 48 similar data of first- 



386 



STEAM POWER PLANT ENGINEERING 



class turbines. A study of these tables will show that the choice must 
be based on other factors than the steam consumption. In a general 
sense, the piston engine is superior to the turbine for high back pres- 
sures, slow rotative speeds and heavy starting torques, while the tur- 
bine has practically superseded the engine for large central station 
units and for auxiliaries requiring high rotative speed. Recent tests 
of the Melville reduction gear (Machinery, Feb., 1910) show exception- 
ally high efficiencies for sizes as large as 6000 kilowatts, and it is not 



80 










Steam Press. -150 Lb 












70 

w 

ft 

W 60 

ft 

u 


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unlikely that the turbine equipped with this device will offset the low 
rotative speed factor of the piston engine. 

If the tests of steam turbines and piston engine could be made at some 
standard initial pressure, back pressure and quality or superheat, then 
a comparison could readily be made, but both types of prime movers 
are designed to give the best results for special operating conditions, 
and any marked departure from these conditions will result in loss of 
economy. It is frequently desired, however, to make a comparison 
between the economy of the different machines, and the following 
methods are in vogue: 

(1) Steam consumption under assumed conditions. 

(2) Heat consumption per horse-power per minute above the ideal 
feed-water temperature. 

(3) Efficiency ratio or ratio of ideal to actual. 



STEAM TURBINES 



387 



Standard Correction Curves. 

This method for comparing engines or turbines or both is best illus- 
trated by a specific example: Suppose it is required to compare the 
full-load performance of a 125-kw. direct-connected piston engine 
with that of a 125-kw. turbo-generator with operating conditions as 
follows: 





Steam Consump- 
tion, Lbs. per 
Kw.-Hour. 


Initial Pres- 
sure, Lbs. 
Absolute. 


Vacuum, 

Inches of 

Hg. 


Superheat, 
Deg. F. 


Engine 

Turbine 


25.0 

22.7 


160 
110 


25.5 
28.0 




125 



Manufacturers of steam turbines have provided correction curves as 
illustrated in Fig. 194d, showing the influence of varying vacuum, 
superheat and pressures on the steam consumption.* From curve 
B, we find that the steam consumption of the turbine should be 



100 



110 



Steam Pressure, Lbs. per Sq. In. Absolute 
120 130 140 150 160 170 180 



190 



200 



30 



t-1 
I 24 

3 



20 































































A 0°F Superheat,165 *"Abs. 

B » » , 28 in Vacuum 

C 165* Absolute, » » 










































A 


^Vi 


c 0li t 


u-Co 






































p^o^C 


W 




























B 


Pr e& 


sure 


Corr 


ectic 


23 


1 




























C 


g"Perhei 




rve 
































rrec 


ion 


?urv 


e 
















TYPICAL CORRECTION CURVES 

FOR 125 K.W. STEAM TURBINE 

FULL LOAD CONDITIONS. 





















































































20 


40 


CO 


80 100 120 140 
Superbeat, Deg. Fab. 


160 


20 


21 


22 


23 24 25 26 

Vacuum, Inches of Mercury 


M 



180 



200 
29 



Fig. 198d. 

decreased 2.5 pounds to give the equivalent at 160 pounds initial 
pressure; from curve A it should be increased 2.5 pounds to give the 
equivalent at 25.5 inches of vacuum, and from curve C it should be 
increased 2.5 pounds to give the equivalent at degree superheat. 
The full-load steam consumption for the turbine under the engine con- 
ditions is therefore 22.7 — 2.5 + 2.5 + 2.5 = 25.2 pounds per kw.-hour. 
* These curves are drawn to a much larger scale than the reproduction given here. 



388 STEAM POWER PLANT ENGINEERING 

The ratio method is also used in this connection, thus: The full-load 
steam consumption at 160 pounds pressure, curve B, Fig. 194d, is 

25 

multiplied by the ratio 7 r=- - to give the equivalent consumption at 110 
Zt .o 

pounds (25 is the steam consumption at 160 pounds and 27.5 the con- 
sumption at 110 pounds). Similarly the correction ratio to change 

25 5 
the consumption at 28 inches of vacuum to 25.5 is -^-, and to correct 

25 

125° F. superheat to 0° F. is — -• 



Summary. 




25 

Pressure correction 7 —— = 0.91 = 

11 .5 


- 9%. 


. 27.5 
Vacuum correction — -^ = 1.10 = 
lb 


10%. 


25 
Superheat correction 7 ^— = 1.11 = 


11%. 



Net correction 12%. 

Corrected steam consumption = 22.7 + 0.12 X 22.7 = 25.4 pounds 
per kw.-hr. 

The ratio method is generally used if the difference between the 
corrected steam consumption and that of the correction curves for 
the same conditions is greater than 5 per cent ("The Steam Turbine," 
Moyer, p. 128). 

This ratio method for correcting steam consumption at full load may 
be used without appreciable error for half to one and one half load and 
is the only practical method for quarter load (Engrg. London, March 2, 
1906). 

Heat Consumption. 

The heat consumption B.T.U. per unit output per minute above 
the ideal feed water temperature may be expressed 

— ^-~- — — • See equation (/5b). 



For the case 


cited above 






Engine, 


25 (1194.1 - 
60 


98) 


= 455 B.T.U. 


Turbine, 


22.7 (1264.2 


-70) 


= 451 B.T.U. 



STEAM TURBINES 



389 



Efficiency Ratio. 

The efficiency ratio, or the extent to which the theoretical possibilities 
are realized, may be expressed 

2545 



Er W {H x - H 2 ) 

For the case cited above 

™ • 2545 

Engine, 

Turbine, 



25(1194.1 - 915) 

2545 

22.7 (1264.2 - 915.3) 



See equation (76b). 

= 0.366. 
= 0.322. 



In the assumed case the turbine is the more economical in heat con- 
sumption, but the engine is the more perfect of the two as far as theo- 
retical possibilities are concerned. 



30 
29 






















































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Load in terms of rated electrical load 



Fig. 199. 



202. First Cost. — Steam turbines, generally speaking, are about 10 
per cent lower in first cost than high-grade compound engines of equiv- 
alent power. The following table gives an idea how the price varies 
with the conditions of operation. The figures are approximate only 
and refer to the cost of the turbine and generator exclusive of auxil- 
iaries. 



390 



STEAM POWER PLANT ENGINEERING 






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STEAM TURBINES 



391 



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392 



STEAM POWER PLANT ENGINEERING 



APPROXIMATE COST OF STEAM TURBINES AND GENERATORS. 

In Dollars ner Kilowatt. Rated Capacity. 





Kilowatts. 




25 

55 
60 


50 

47 
51 


75 

42 
46 


100 

38 
43 


200 

32 
36 


300 


500 


1000 


2000 


4000 


6000 


Direct current: 
Non-condensing .... 


32 
36 

35 
35 






Condensing 












Alternating current: 
25 cycles 


32 
30 


28 
27 


25 
25 


21 
21 


20 


60 cycles 












20 

















203. Cost of Operation. — Data pertaining to the cost of operating 
steam-turbine and reciprocating engine plants and combinations of 
both will be found in Chapter XVII. The following table, contributed 
by H. G. Stott, Superintendent Motive Power of the Interborough 
Rapid Transit Company, New York, gives an excellent comparison of 
the relative maintenance and operating costs (March, 1910) of the 
three types of steam power plants as applied to large central stations 
for electric street railways. 

RELATIVE COSTS PER KILOWATT-HOUR. DISTRIBUTION OF 
MAINTENANCE AND OPERATION. 





Reciprocating 
Steam 
Plant. 


Steam 

Turbine 

Plant. 


Reciprocating 
Engines and Low- 
pressure Steam 
Turbines. 


Maintenance. 

1. Engine room, mechanical 

2. Boiler or producer room 

3. Coal and ash handling apparatus. . . . 

4. Electrical apparatus 

Operation. 

5. Coal 

6. Water 


2.59 

4.65 
0.58 
1.13 

61.70 
7.20 
6.75 
7.20 
2.28 
1.07 
2.54 
1.78 
0.30 
0.17 


0.51 
4.33 
0.54 
1.13 

55.53 
0.65 
1.36 
6.74 
2.13 
0.95 
2.54 
0.35 
0.30 
0.17 


1.55 
3.55 
0.44 
1.13 

46.48 
0.61 


7. Engine room labor 


4.06 


8. Boiler or producer room labor 

9. Coal and ash handling labor 


5.50 
1.75 
0.81 


11 Electrical labor 


2.54 


12 Engine room lubrication 


1.02 


13. Engine room waste, etc 


0.30 
0.17 






Relative operating cost, per cent. . . . 

Relative investment, per cent 

Probable average cost per kw 


100.00 
100.00 
125.00 
11% 


77.23 
75.00 
93.75 
11% 


69.91 

80.00 

100.00 

11% 





For steam turbine plants larger than 60,000 kw. the cost per kilowatt 
may be reduced to $75.00. 






STEAM TURBINES 



393 



204. Influence of Superheat. — The use of superheated steam in- 
creases the economy of the reciprocating engine about 1 per cent for 
every 10 to 20 degrees of superheat, depending upon the conditions of 
operation, the gain being due mainly to the reduction of cylinder con- 
densation. Cylinder condensation is reduced not only because of the 
excess heat available for the evaporation of moisture but also because 
superheated steam has a lower conductivity than wet steam, and less 
heat is given up to the cylinder walls for the same difference of tem- 
perature. In the steam turbine this difference of temperature is much 
smaller, since high- and low-pressure steam do not alternately come in 



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vj 28 inches Vacuum 
Tested by Messrs. Dean & Main Eng'rs. 


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180. 



contact with the same surface as is the case with the reciprocating 
engine, and the time of contact is considerably less, due to the com- 
paratively high velocities. With a well-lagged casing, therefore, the 
condensation due to this cause is insignificant compared with that of 
the reciprocating engine, and the beneficial effect of superheat is much 
more pronounced. Friction of the steam, which in the reciprocating 
engine is negligible, and which may be a source of considerable loss in 
the turbine, is greatly reduced by the use of superheated steam, as is 
also the " windage " loss due to the rapid revolution of the wheels. 

The problem of cylinder lubrication is sometimes a difficult one in 
steam engines using a high degree of superheat, and trouble is fre- 
quently experienced due to the unequal expansion of the metal. In 



394 



STEAM POWER PLANT ENGINEERING 



the steam turbine the latter difficulty is not so pronounced and no 
internal lubrication is necessary, hence higher degrees of superheat 
are permissible. For maximum economy the steam at the end of 
expansion should be free from moisture. Assuming purely adiabatic 
expansion, the steam in expanding from 165 pounds to 1 pound abso- 
lute would have to be superheated about 500 degrees F., giving the 
steam an actual temperature of 800 degrees F. A study of some 100 
tests made in this country gives about 250 degrees superheat as a maxi- 





















































































































































RELATI 


ON- VACUUM TO ECONOMY 
3 K.W. Parsons Turbine 








36 






















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Fig. 201. 



28__30 



mum and 100 degrees to 150 degrees F. as an average. In Europe 
reciprocating engines are operating with superheat as high as 350 
degrees F. and turbines 300 degrees F. The additional fixed and 
operating costs of superheating must be considered in determining the 
net gain, since the decrease in steam consumption does not represent 
the actual saving. With pressures of 175 pounds gauge or less, and 
not to exceed 200 degrees F. superheat, the net gain has in most cases 
proved a substantial one. With higher temperatures and pressures the 
cost of maintaining the superheat may increase more rapidly than the 
saving in steam consumption, until a limit is reached beyond which no 






STEAM TURBINES 



395 



advantage is gained. The relation between superheat and steam con- 
sumption for a 400-kilowatt Westinghouse-Parsons turbine is illustrated 
in Fig. 200. Fig. 202 gives a similar comparison for a 1500-kilowatt 
turbine. (J. R. Bibbins, Power, January, 1905.) 

205. Influence of High Vacua. — The possible economy of the recip- 
rocating engine is greatly restricted by its limited range of expansion. 
Cylinders cannot be profitably designed to accommodate the rapid 
increase in the volume of steam when expanded to very low pressures. 
For example, the specific volume of 1 pound of steam under a vacuum 
of 29 inches (referred to a 30-inch barometer) is about 650 cubic feet, 
or nearly double its volume under a vacuum of 28 inches. Usually 
the exhaust is opened at a pressure of 6 or 8 pounds absolute and 
consequently a large proportion of the available energy is lost. The 
lower vacuum in the exhaust pipe, therefore, serves only to diminish 
the back pressure and does not affect the completeness of expansion. 
Even if it were practical to expand to 1 pound absolute, the increased 
condensation in the reciprocating engine would offset any gain due to 
expansion unless the steam were highly superheated. A study of a 
number of tests of reciprocating engines shows a slight improvement 
due to increasing the vacuum beyond 26 inches. Tests of steam tur- 
bines show a decrease of 3 to 4 per cent in steam consumption for each 
inch increase of vacuum between 25 and 29 inches, for with a well- 
lagged casing cylinder conden- _ 
sation is practically absent, 
since the high- and low-temper- 
ature steam do not alternately 
come in contact with the metal- 
lic surfaces as is the case with the 
reciprocating engine (Figs. 201 
and 201a). Fig. 230 shows a 
relation between the power con- 
sumption of the auxiliaries and 
the total output of the station 
at different loads for a Parsons 
steam turbine installation and 
Fig. 231 shows a similar relation 
for the 2000-kilowatt Curtis tur- 
bine. The power consumption 
in the latter case is higher on account of the high temperature of cool- 
ing water. Table 55 gives the power required for the auxiliaries in a 
number of stations. A high vacuum may be limited by the initial tem- 
perature of the cooling water. The difference in temperature between 



















I Mill II 1 1 1 1 1 1 












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1 1 1 1 1 II 1 1 1 1 1 1 1 1 1 












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a subsequent test under same 
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27 .2 .4 .6 .8 28 



396 



STEAM POWER PLANT ENGINEERING 



16 
15.5 
15 
04.5 

14 
H'3.5 
03 
12.5 
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Efi. ect of Vacuum and Superheat 


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1500 K.Vf . .Turbine ,Full Load 


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26 27 

Vacuum Inches 



20 40 60 80 100 120 140 
Superheat Deg.E. 



Fig. 202. 



inlet and discharge should be greater than 10 degrees, since otherwise 
the amount of circulating water per pound of steam becomes excessive 
and increases the work of the pumps. For example, the temperature 
of steam corresponding to a vacuum of 28 inches or 1 pound absolute 

is 102 degrees, and with cooling 
water at 75 degrees F., and the 
discharge at 95 degrees F., the 
theoretical ratio of cooling water 
to steam necessary to maintain 
this vacuum will be about 50 
and the actual nearer 70. From 
Table 50 it will be seen that 
a 28-inch vacuum referred to a 
30-inch barometer is obtained 
with an average ratio of 50 
pounds of cooling water at 70 
degrees F. per pound of steam. The cost of high- vacuum apparatus is 
not proportional to the vacuum, but increases much more rapidly, as 
shown in Fig. 232. These estimates show averages and not specific costs. 
Fig. 202 shows the effect of superheat and vacuum on the economy 
of a 1500-kilowatt Westinghouse-Parsons turbine. Figs. 200 to 202 are 
taken from " Steam Power Plants," by J. R. Bibbins, as published in 
Power, January, 1905. 

Reciprocating Engine vs. Turbine: Power, April, 1904, p. 232, May, 1904, p. 298; 
Engr. U. S., Nov. 1, 1905, p. 711; Elec. World, April 2, 1904, p. 651; Eng. Mag., 
Sept., 1905, p. 935; Power, Feb., 1906, p. 83; Elec. Age, June, 1905, p. 478; Elec. 
Rev., Dec. 23, 1904. 

Steam Turbine Design: Eng. Rec, July 22, 1905, p. 101 ; St. Ry. Jour., Dec. 20, 
1902, p. 988; Engr., Lond., Jan. 8, 1904, p. 34, May 13, 1904, p. 481, Dec. 27, 1907; 
Electrician, Lond., March 24, 1905; Power, Dec, 1905; Mech. Eng., Feb. 7, 1908; 
Engineering, Dec. 13, 1907. 

Theory and Design of Steam Turbines: Engr. U. S., Dec. 16, 1907, p. 1126 (serial), 
March 15, p. 201; Revue de Mecanique, Oct. 31, 1907; Engr., Lond., Oct. 4, 1907; 
Eng. Rev., May, 1904; Mech. Engr., Oct. 28, 1905; Eng. Rec, July 22, 1905, p. 101, 
May 7, 1904, p. 581. 

Modern Steam Turbine Plants: Power, Dec, 1906, p. 717, Dec, 1907; Engr. U.S., 
Nov. 15, 1906, p. 733, March 15, 1907, p. 304; St. Ry. Jour., Oct. 19, 1907; Elec. 
World, July 22, 1905, Feb. 15, 1908; Eng. Rec, March 4, 19Q5. 

Tests of Westinghouse-Parsons Turbines: Engr. U.S., Dec. 1, 1904, p. 802; St. Ry. 
Jour., Dec. 19, 1903, p. 1063; Power, March, 1904, p. 127, April, 1904, p. 239, Aug., 
1905, p. 466; Elec. World, Sept. 6, 1902, p. 360; Eng. Rec, July 29, 1905, p. 134. 

Governing Steam Turbines: Harvard Engineering Jour., 1908. 

A Recent Comparison of Turbines and Engines: Eng. Rec, Feb. 19, 1910. 

Internal Losses of a Steam Turbine: Power & Engr., Aug. 24, 1909. 

The Principles of Steam-turbine Buckets: Power, Mar. 17, 1908. 



CHAPTER XI. 

CONDENSERS. 

206. General. — A pound of dry steam at atmospheric pressure (30 
inches mercury) occupies a volume of 26.8 cubic feet. Suppose these 
26.8 cubic feet of steam were contained in a closed vessel, and that the 
steam was subsequently condensed and its temperature lowered by 
suitable means to say 110 degrees F. The condensed steam would 
occupy only about T 7V0 °f its original volume, and the pressure would 
fall to 2.6 inches of mercury, the latter pressure being due to the ten- 
sion of the aqueous vapor at the given temperature. That is to say, 
the best vacuum theoretically attainable under the given conditions 
would be 30 — 2.6 = 27.4 inches. The lower the temperature to 
which the condensed steam is reduced the more nearly perfect will be 
the vacuum attained. 

If air is mixed with the steam the vacuum will be still more imperfect. 
Thus, suppose the vessel to contain one pound of steam and one-tenth of 
a pound of air under atmospheric pressure. The volume of the closed 
vessel in this case must be 26.8 + 1.69 = 28.49 cubic feet. 

After the steam has been condensed and its temperature reduced to 
110 degrees F. the remaining pressure will be due to the aqueous vapor 
tension plus the pressure due to the air, or 2.6 + 1.51 = 4.11 inches 
mercury, and the maximum vacuum attainable will be 25.89 inches. 
In practice air is always present in exhaust steam. A condenser is a 
device in which the process of condensation and subsequent removal 
of the air and condensed steam is continuous, the degree of vacuum 
obtained depending upon the tightness of valves and joints, the quan- 
tity of entrained air, and the temperature to which the condensed steam 
is reduced.* 

The degree of vacuum may be expressed in different ways. (1) Excess 
of the atmospheric pressure over the observed vacuum. For example, 
a 26-inch vacuum implies that the pressure of the atmosphere is 26 
inches of mercury above the pressure in the condenser. (2) Per cent 
of vacuum, by which is meant the ratio of the observed vacuum to 
the atmospheric pressure. Thus with the barometer standing at 30 
inches a vacuum of 26 inches may be expressed as 100 X ff = 86.6 
per cent vacuum. This method of expression gives an idea of the 

* See " The Influence of Air on Vacuum in Surface Condensers," Engng., April 
17, 1908; Power, Feb. 2, 1909, p. 235. 

397 



398 



STEAM POWER PLANT ENGINEERING 



efficiency of the condensing system. For example, the degree of 
vacuum indicated by 26 inches would be 93 per cent with a barometric 
pressure of 28 inches but only 84 per cent when the barometer reads 31 
inches. (3) Absolute pressure. Thus a 26-inch vacuum referred to a 
30-inch barometer would be indicated as a pressure of 30—26 = 4 inches 
absolute, or 1.9 pounds per square inch. 

207. Function of the Condenser. — The function of a condenser in 
connection with a steam engine or turbine is primarily the reduction of 
back pressure, though in some instances, notably in marine work, the 
recovery of the condensed steam may be of equal importance. The 
advantages to be gained by decreasing back pressure may be most 
readily illustrated by the following example: A non-condensing engine 
taking steam at a pressure of 100 pounds absolute and cutting off at 
one-quarter stroke will have, theoretically, a mean effective pressure 
on the piston of 44.6 pounds per square inch, the back pressure being 
14.7 pounds per square inch absolute. If the engine exhausts into a 
condenser against a 26-inch vacuum (1.7 pounds absolute) the mean 
effective pressure will be increased to 44.6 + (14.7 — 1.7) = 57.6 
pounds per square inch, resulting in a gain in power which may be 
expressed 

H.P. = IjzA/L , 
33,000 

in which 



(82) 



H.P. = horse power gained. 
P r = reduction in back pressure, pounds per square inch. 
A = area of the piston in square inches. 
S = piston speed in feet per minute. 



If P = mean effective pressure on the piston when running non-con- 
densing, the percentage of increase of power may be expressed 



Per cent = 100 



Pr 



(83) 



In the above example the percentage of power gained would be 

13 



100 



44.6 



29.2 per cent. 



The actual gain due to the use of the condenser would be much 
less than this, depending upon the type of engine and conditions of 
operation, as shown in the results of engine performances outlined in 
Chapter X. 



CONDENSERS 



399 



TABLE 49. 
PRESSURE OF AQUEOUS VAPOR IN INCHES OF MERCURY FOR EACH DEGREE F. 

(Marks and Davis.) 





0° 


1° 


2° 


3° 


4° 


5° 


6° 


-0 


8° 


9° 


30° 






.180 
.268 
.390 
.560 
.790 
1.10 
1.51 
2.04 
2.74 
3.63 
4.76 
6.18 


.188 
.278 
.405 
.580 
.817 
1.13 
1.55 
2.11 
2.82 
3.74 
4.89 
6.34 


.195 
.289 
.420 
.601 
.845 
1.17 
1.60 
2.17 
2.90 
3.84 
5.02 
6.51 


.203 
.300 
.436 
.622 
.873 
1.21 
1.65 
2.24 
2.99 
3.95 
5.16 
6.67 


.212 
.312 
.452 
.644 
.903 
1.25 
1.71 
2.30 
3.07 
4.06 
5.29 
6.84 


.220 
.324 
.468 
.667 
.964 
1.30 
1.76 
2.37 
3.16 
4.17 
5.43 
7.02 


.229 
.336 
.486 
.690 
.996 
1.33 
1.81 
2.44 
3.25 
4.28 
5.58 
7.20 


.238 


40° 


.248 
.362 
.522 
.739 
1.03 
1.42 
1.93 
2.60 
3.44 
4.52 
5.88 


.257 
.376 
.541 

.764 
1.06 
1.46 
1.98 
2.66 
3.53 
4.64 
6.03 


.349 


50° 

60° 

70° 


.503 

.714 
1.03 


80° 


1.37 


90° 

100° 

110° 


1.87 
2.51 
3.34 


120° 

130° 


4.40 
5.73 


140° 


7.38 







With steam turbines the advantage gained by reduction of back 
pressure is more marked than with the reciprocating engine, though 
theoretically the same for the same range of expansion. Initial con- 
densation, leakage past valves, and other sources of loss prevent a 
reciprocating engine from benefiting from a good vacuum to the same 
extent as a turbine. (See paragraph 205.) 

Referring again to the example given above, if the steam is cut off at 
about one-sixth stroke, the work done when running condensing will be 
the same as when running non-condensing and cutting off at one-quarter. 
Theoretically the steam consumption will be decreased nearly in pro- 
portion to the reduction in cut-off. Generally speaking, a condensing 
engine will require from 20 to 30 per cent less steam for a given power 
than a non-condensing engine. (See results of engine tests, paragraph 
179.) This decrease in steam consumption is only an apparent one. If 
steam is used by the auxiliaries in creating the vacuum, the amount must 
be added to that consumed by the engine, unless the steam exhausted 
by the former is utilized to warm the feed water, in which case only the 
difference between the heat entering the auxiliaries and that returned 
to the heater should be charged against the engine. The power neces- 
sary to operate the condenser auxiliaries varies from one to six per cent 
of the main engine power, depending upon the type and conditions of 
operation. (See paragraph 228.) 

In power plants where the exhaust steam is not used for heating or 
manufacturing purposes, the engines are almost invariably operated 
condensing, provided there is an abundant supply of cooling water. 
Even if the water supply is limited, it is often found to be economical to 



400 



STEAM POWER PLANT ENGINEERING 



use some artificial cooling device, notwithstanding the high first cost 
and cost of operation of the latter. 

Some of the considerations affecting the propriety of running con- 
densing and the choice of condensing systems are taken up in para- 
graphs 230 and 231. 

The Law of Condensation of Steam : Pro. Inst, of Civil Engrs., Nov. 30, 1897; 
Engr., Lond., Dec. 17, 1897, p. 609. Relation of Pressure and Temperature in Con- 
densers : Power, April, 1902, p. 28, June, 1902, p. 30. The Measurement of Vacuum : 
Engr., Lond., April 21, 1905. Experiments on Condensation of Steam : Engr., Oct. 
15, 1897, p. 481. Condensation Fallacies and Facts: Machinery, Sept., 1904, p. 38; 
Elec. Review, June 17, 1904. Condenser Pressure, a Neglected Point in Steam 
Condensers : Mech. Engr., Sept. 30, p. 484. 

Advantages of Condensing : Amer. Elecn., Sept., 1904, p. 469. Condensers for 
Steam Engines and Turbines (F. Foster) : Mech. Engr., Oct. 28, 1905, pp. 637, 655. 
Condensers, to What Extent Should They be Used : Power, Nov., 1899. The Value 
of the Condenser : Power, Oct., 1902. The Effects of Vacuum on Steam Engine 
Economy : Eng. Mag., June, 1905. Importance of Condensers : Power, May, 1897; 
Evaporating, Condensing, and Cooling Apparatus (E. Hausbrand), published by 
Scott, Greenwood & Co., London. 

The Use of Condensers : Electrician, Lond., Oct. 12, 1900. 



208. Classification of Condensers. — The following is a classification 
of a few well-known steam engine condensers : 



1. Jet condensers. 



2. Surface condensers. 



Parallel current (a). 



Counter current (6). 



Worthington. 

Blake. 

Deane. 

Baragwanath. 

Bulkley. 
Ejector (3) Schutte. 

Weiss. 
Barometric. 



Ordinary (1) 
Siphon (2).., 



Alberger. 
Tomlinson. 

Single-flow Baragwanath. 

Double-flow Wheeler. 

Multi-flow Wainwright. 

Forced draft Fouche. 

i Natural draft Pennell. 

Evaporative (c) Ledward. 



Water cooled (a) 
Air cooled (6) . . . . 



Condensers may be divided into two general groups: 

1. Jet condensers, in which the steam and cooling water mingle and 
the steam is condensed by direct contact, Figs. 203 to 211. 

2. Surface condensers, in which the steam and cooling medium are 
in separate chambers and the heat is abstracted from the steam by con- 
duction, Figs. 212 to 215. 



CONDENSERS 401 

Jet condensers may be further grouped into two classes, according to 
the direction of flow of the air and cooling water: 

(a) Parallel-current condensers, in which the condensed steam, cool- 
ing water, and air flow in the same direction, collect at the bottom of 
the condenser chamber, and are exhausted by a suitable pump, 
Fig. 203. 

(b) Counter-current condensers, in which the cooling water and con- 
densed steam flow from the bottom of the chamber, usually by gravity, 
while the air is drawn off at the top, Fig. 229b. 

Parallel-current condensers may be subdivided into three classes: 

(1) Standard condensers, in which the cooling water, condensed 
steam, and air are exhausted by a vacuum pump, Figs. 203 and 
209. 

(2) Siphon condensers, in which the cooling water, condensed steam, 
and air are exhausted by a barometric column, Fig. 205. 

(3) Ejector condensers, in which the condensed steam and air 
are exhausted by the cooling water, on the ejector principle, Fig. 
206. 

Surface condensers may be classified according to the nature of the 
cooling medium as 

(a) Water-cooled condensers, Fig. 212. 

(b) Air-cooled condensers, Fig. 217. 

(c) Evaporative condensers, in which the condensation of the steam is 
brought about by the evaporation of a fine stream of water trickling on 
the outside of the tubes. 

309. Ordinary Jet Condensers. — Fig. 203 shows a section through a 
Worthington jet condenser, illustrating the parallel-current principle. 
When the pump is started a partial vacuum is created in the suction 
chamber above the valves H, H in the cone F. As soon as sufficient 
air has been exhausted, cooling water enters at B with a velocity depend- 
ing upon the degree of vacuum in chamber F and the suction head, 
and is divided into a fine spray by the adjustable serrated cone D. 
The spray mingles with the exhaust steam entering at A and both 
move downwards with diverse velocities. The steam gives up its heat 
to the water and condenses. The velocity of the steam diminishes in its 
downward path to zero, while the velocity of the water increases accord- 
ing to the laws of falling bodies. The condensed steam, cooling water, 
and air collect at the lower part of the condenser and are exhausted by 



402 



STEAM POWER PLANT ENGINEERING 



the wet air pump G, from which they are forced through opening J to 
the hot well. The vacuum in chamber F will depend upon the vapor 
tension of the warm water in the bottom of the well, the amount of 
air carried along by the cooling water and steam, and the tightness of 




Fig. 203. Worthington Independent Jet Condenser. 



valves and joints. In case the water accumulates in the condenser 
cone F, either by reason of an increased supply or by a sluggishness or 
even stoppage of the pump, the condensing surface is reduced to a mini- 
mum, as soon as the level of the water reaches the spray pipe and the 
spray becomes submerged, and only a small annular surface of water is 
exposed to the exhaust steam. The vacuum is immediately broken, 
and the exhaust steam escapes by blowing through the injection pipe 






CONDENSERS 



403 



and through the valves of the pump and out the discharge pipe at J, 
forcing the water ahead of it; consequently flooding of the steam 
cylinder cannot occur. In starting up the condenser a partial vacuum 
for inducing a flow of injection water into the condenser chamber may 
be created by the pump if the suction lift is not too great. Many 
engineers, however, prefer to install a small forced injection or priming 
pipe the function of which is to condense sufficient steam to produce 
the necessary partial vacuum. Fig. 222 shows such an installation. 




AIR PUA 

Fig. 204. Section through a Blake Jet Condenser. 



Fig. 204 shows a section through the condensing chamber and air 
pump of a Blake vertical jet condenser with an automatic vacuum- 
breaking device. The injection water enters at opening marked 
" injection " and flows through the adjustable " spray " nozzle in a fine 
spray, at an angle of about 45 degrees, and impinges on the conical sides 
of the upper condenser chamber. The spray falls from the sides to the 
projecting ledges shown in the illustration. The ledges prevent the 



404 STEAM POWER PLANT ENGINEERING 

spray from falling directly to the bottom of the chamber and insure an 
efficient mingling of steam and cooling water. A perforated copper 
plate is substituted for the shelves when the force of the injection water 
is not sufficient to produce spray. The circulating water and con- 
densed steam together with the non-condensable gases are drawn off at 
the bottom of the chamber. The vacuum-breaking device is shown 
at the right of the figure. When the rising water reaches the level of 
the float chamber, as in the case of an accidental stoppage of the air 
pumps, the float is raised and forces a check valve from its seat and 
allows an inrush of air to break the vacuum, thus preventing further 
suction of water into the condenser and consequent flooding of the 
engine. A is the forced injection or " priming " inlet used in start- 
ing up when the suction lift is considerable. 

Condenser Types and Applications: Power, June, 1906, p. 44; Engr. U.S., Jan., 
1906, pp. 55-66. Jet Condensers (McBride): Trans. A.S.M.E., Vol. 12, p. 187; 
American Machinist, March 7, 1895, p. 185; Engr. U.S., Jan., 1906, p. 61; Whitham's 
Steam Engine Design, p. 294; Seaton's Manual of the Marine Engine, Chapter XI. 

210. Condensing Water, Jet Condensers. — In a jet condenser the 
cooling water and exhaust steam mingle, and the degree of vacuum is 
a function of the final or discharge temperature; thus the quantity of 
cooling water required depends upon its initial temperature, the tem- 
perature of the discharge water, and the total heat in the steam entering 
the condenser. If the steam in the low-pressure cylinder at release is 
dry and saturated, the heat entering the condenser will correspond 
to the total heat in steam at release pressure, but it usually contains 
considerable moisture, part of which is reevaporated when the exhaust 
valve opens to the condenser; however, it is sufficiently accurate for 
all practical purposes to assume the exhaust steam entering the con- 
denser to be dry and saturated and its heat to correspond to the pres- 
sure in the condenser. 

Let A = total heat of steam at condenser pressure above 32 degrees. 
T 2 = temperature of the discharge water. 
T = initial temperature of the cooling water. 
W = weight of cooling water in pounds necessary to condense 
and cool one pound of steam to the required discharge 
temperature. 

Then W = X ~ r »t 32 ' ( 84 ) 

T 2 - T 

Example: How many pounds of cooling water are necessary to con- 
dense one pound of steam under the following conditions: Barometer 
29.92; vacuum 26 inches; temperature of injection water 60 degrees F. 



CONDENSERS 405 

The temperature of aqueous vapor corresponding to an absolute 
pressure of 29.92 — 26 = 3.92 inches of mercury is 125 degrees F. (See 
Table 49.) The discharge temperature, however, must be less than this, 
as the pressure in the condenser is due not only to the aqueous vapor but 
to that of the air carried over with the circulating water and the con- 
densed steam. In a condenser of this type the discharge temperature 
will be from 10 degrees to 15 degrees lower than that corresponding to 
the vacuum as recorded by the gauge. In this case assume it to be 15 
degrees lower, i.e., T 2 = 125 — 15 = 110 degrees. 

The total heat corresponding to a pressure of 3.92 inches of mercury 
is 1120 B.T.U. above 32 degrees (see steam tables); T = 60 degrees; 
T 2 = 110 degrees. 

w= 1120-110 + 32 
110-60 

Evidently the higher the temperature of the discharge water the less 
will be the quantity of cooling water required, and consequently the 
smaller the weight of air introduced into the condenser; but the warmer 
the discharge water the greater will be the vapor tension and the lower 
the degree of vacuum. For reciprocating engines a hot-well tempera- 
ture between 110 degrees and 130 degrees F. is average practice; with 
turbines the temperature ranges between 80 degrees and 100 degrees F. 
On account of the inefficient heat absorption in practical installations, 
from 5 per cent to 15 per cent is added to the theoretical weight of cool- 
ing water as determined from equation (84). Table 50 has been calcu- 
lated from equation (84). 

Cooling Water for Condensers: Am. Mach., May 18, 1905, p. 656; Evaporation 
and Condensing Apparatus, Hausbrand, pp. 227, 240, 301, 318; Steam Power Plants, 
Meyer, p. 106. 

Wet-Air Pump, Jet Condensers. (See paragraph 285.) 

Circulating Pumps. (See paragraph 297.) 

211. Effect of Aqueous Vapor upon the Degree of Vacuum. — The 

futility of attempting to better the vacuum by exhausting the vapor is 
best illustrated by a specific problem. 

Required the volume of aqueous vapor to be withdrawn per hour 
from a condenser operating under the following conditions, in order 
that the vacuum may be increased one pound per square inch : Tem- 
perature of discharge water 125 degrees; corresponding vapor tension 
4 inches of mercury; barometer 30 inches; relative vacuum 26 inches; 
horse power, 100; steam consumption 20 pounds per horse-power hour; 
cooling water 25 pounds per pound of steam condensed. 

100 X 20 X 25 = 50,000 pounds of cooling water per hour. 
= 833 pounds of cooling water per minute. 



TABLE 50. 
RATIO, BY WEIGHT, OF COOLING WATER TO STEAM CONDENSED (THEORETICAL). 











(Barometer 29.92.) 












Vacuum 24". 
Temperature of Steam 141°. 


Temp. 
of In- 
jection. 


Vacuum 25". 
Temperature of Steam 134°. 


Temp, 
of In- 
jection. 


Temperature of Hot Well. 


Temperature of Hot Well. 




110 


115 


120 


125 


130 


105 


110 


115 


120 


125 


40 
50 
60 
70 
80 
90 


15.0 

17.5 
21.0 
26.2 
35.0 
52.4 


13.9 
16.0 
18.9 
23.2 
29.8 
49.7 


12.9 

14.8 
17.3 
20.7 
25.9 
34.6 


12.1 
13.7 
15.8 
18.7 
23.0 
29.5 


11.4 
12.8 
14.6 
17.1 
20.5 
25.6 


40 
50 
60 

70 
80 
90 


16.1 
19.0 
23.2 
30.0 
42.0 
70.0 


14.9 
17.4 
20.9 
26.1 
34.8 
52.1 


13.8 
16.0 
18.9 
23.0 
29.6 
41.5 


12.9 

14.8 
17.2 
20.7 
25.9 
34.5 


12.1 
13.7 
15.8 
18.7 
22 8 
29.4 




Vacuum 26 ,; '. 
Temperature of Steam 125°. 


Temp. 
of In- 
jection. 


Vacuum 27". 
Temperature of Steam 114°. 


Temp, 
of In- 
jection. 


Temperature of Hot Well. 


Temperature of Hot Well. 




100 


105 


110 


115 




90 


95 


100 


105 




40 
50 
60 
70 

80 


17.5 
21.0 
26.3 
35.0 
57.6 


16.1 

19.0 
23.2 
30.0 
42.0 


14.8 
17.4 
20.9 
26.0 
34.7 


13.8 
16.0 
18.8 
23.0 
29.6 




40 
50 
60 
70 

80 


21.2 

26.5 
35.3 
52.9 


19.1 
23.4 
30.1 
42.1 


17.4 
20.9 
26.2 
34.9 
52.3 


16.0 
19.0 
23.2 

29.8 
41.5 













Temp. 
of In- 
jection. 


Vacuum 27.5". 
Temperature of Steam 108°. 


Temp, 
of In- 
jection. 


Vacuum 28". 
Temperature of Steam 100°. 


Temperature of Hot Well. 


Temperature of Hot Well. 




80 


85 


90 


95 




75 


80 


85 


90 




40 
50 
60 
70 


26.6 
35.6 
52.3 


23.6 
30.3 
42.5 
70.8 


21.1 
26.4 
35.2 
52.8 


19.1 
23.4 
30.0 
42.0 




40 
50 
60 
70 


30.5 
42.7 
71.2 


26.6 
35.5 
53.2 


23.5 
30.2 
42.3 
70.6 


21.1 
26.3 
35.1 
52.7 













Temp, 
of In- 
jection. 


Vacuum 28.5". 
Temperature of Steam 90 


o 


Temp, 
of In- 
jection. 


Vacuum 29". 
Temperature of Steam 77 


3 


Temperature of Hot Well. 


Temperature of Hot Well. 




60 


65 


70 


75 




55 


60 


65 


67 




35 
40 
45 
50 


42.2 
54.0 
72.0 


35.8 
43.0 
53.5 
72.0 


30.6 
35.6 
42.8 
53.5 


29.2 
33.4 
38.8 
46.6 




35 
40 
45 
50 


52.0 
69.3 


43.0 
54.0 
71.5 


35.8 
43.0 
54.0 
72.0 


33.4 

38.4 
47.0 
61.0 













CONDENSERS 407 

Now to increase the vacuum one pound per square inch, approxi- 
mately 2 inches of mercury, the temperature of the water must be 
lowered to 102 degrees F., that is, 833 (125-102)= 19,159 B.T.U. 

19 159 

must be abstracted from the water in one minute, or — = 18.6 

pounds of water to be evaporated per minute. (1030 = average heat 
of vaporization of water under 26 to 28 inches of vacuum.) Now, one 
pound of vapor at 102 to 125 degrees F. has an average volume of 270 
cubic feet. 

Therefore 18.6 X 270 = 5022 cubic feet of vapor must be exhausted 
per minute to increase the vacuum from 26 to 28 inches, which is man- 
ifestly impracticable. 

212. Injection Orifice. — The velocity of water entering a jet con- 
denser, neglecting friction, may be determined from the formula 

V = y/Tgh, (85) 

where 

V = velocity of the water in feet per second. 
g = acceleration of gravity = 32.2. 
h = total head in feet. 
If p = pressure below the atmosphere in pounds per square inch, 
h x = distance in feet between the source of supply and the 
injection orifice, 

then h = 2.3 p ± h v (86) 

and equation (85) may be written 



V= 8.025 V2.Sp± h v (87) 

If the supply is under pressure, h x is positive; if under suction, it is 
negative. 

Example: What is the theoretical velocity of water entering a con- 
denser with 26-inch vacuum (referred to 30-inch barometer); suction 
head 8 feet 

Here p = pressure in pounds per square inch, corresponding to 
26 inches of mercury = 12.8 pounds per square inch. 

h t = 8. 

V = 8.025 V2.3 X 12.8-8 
= 37.1 feet per second 
= 2226 feet per minute. 

In proportioning the injection orifice in practice the maximum 
velocity of flow is assumed to be between 1500 and 1800 feet per minute, 



408 



STEAM POWER PLANT ENGINEERING 



or, approximately, area of injection orifice in square inches = weight of 
injection water in pounds -r- 650 to 780. (" Manual of Marine Engineer- 
ing," Seaton, p. 204.) A rough rule gives area of orifice = area of low 
pressure piston in square inches -^ 250. (Seaton, p. 204.) 

213. Volume of the Condenser Chamber. — According to Thurs- 
ton the volume of a jet condenser should be from one-fourth to one- 
half that of the low-pressure engine cylinder. ( " Steam Engine 
Manual," Thurston, II, 127.) 

According to Hutton the volume should not be less than that of the aif 
pump and should approximate three- 
fourths of that of the engine cylinder 
in communication with it. 

214. Injection and Discharge Pipes. 
— In practice the diameter of the injec- 
tion pipe is based on a velocity of 400 
to 600 feet per minute and that of the 
discharge pipe of 200 to 400 feet per 
minute; the lower figures for pipes 
under 8 inches in diameter, the upper 
range for larger diameters. 

(Atmospheric relief valves. — See 
paragraph 351.) 

215. Siphon Condensers. — Fig. 205 
shows a section through a Baragwa- 
nath siphon condenser, illustrating 
the principles of a parallel-current baro- 
metric condenser. The cooling water 
enters the side of the condenser cham- 
ber at A and passes downward in a 
thin annular sheet around the hollow 
cone D. The exhaust steam enters at 
B and is given a downward direction 
by the goose neck C. It flows through 
the nozzle D and is condensed within the hollow cone of moving water, 
the combined mass including the entrained air discharging through the 
contracted throat E at high velocity into the tail pipe F. The water 
column in the tail pipe must be enough to overcome the pressure of the 
atmosphere; i.e., it should be 34 feet or more above the surface of the 
hot well, otherwise water would rise within this pipe to a height corre- 
sponding to that of the barometer, which is approximately 34 feet for a 
barometric pressure of 30 inches of mercury. This is not strictly true 
when the condenser is in full operation, as the injector effect of the 




I ^ u 

I pi 


[TO— 


i Ij^ 


i 





Fig. 205. Baragwanath Siphon 
Condenser. 






CONDENSERS 



409 



moving mass is sufficient to overcome several pounds pressure, and the 
tail pipe may be less than 34 feet, but to provide against any possibility 
of the water being drawn into the cylinder of the engine the length is 
made greater than 34 feet. The spray cone D is adjustable and admits 
of close regulation of the water supply without changing the annular 
form of the stream. The condensing water may be supplied under 
pressure or under suction. For lifts not greater than 15 feet no supply 
pump is necessary, the water being raised by the siphon action of the 
condenser. This condenser requires the same amount of cooling water 
per pound of steam as the standard jet condenser, and is capable of 
maintaining a vacuum of from 24 to 25 inches. A vacuum of 28£ inches 
has been recorded for a condenser of this general type. (Trans. 
A.S.M.E., 26-388.) An atmospheric relief valve G is provided in 
case the vacuum fails from any cause, which will permit the steam to 
escape to the atmosphere. 

The above type of condenser is adapted to very muddy cooling water, 
since no filtration is necessary beyond the removal of such solid matter 
as may clog up the annular space H. 

In the Armour Glue Works at Chicago condensers of this type are 
successfully maintaining a 90 per cent vacuum with cooling water at 
60 degrees F., and the circulating water is practically liquid mud. 

Siphon Condensers, Discussion: Trans. A.S.M.E., Vol. 26, p. 388. Siphon Con- 
densers: Electrical World, June, 1897, p. 818; Hutton, The Mechanical Engineering 
of Power Plants, p. 106; Engr. U.S., Jan., 1906. 

216. Size of Siphon Condensers. — The size of siphon is indicated by 
the diameter of the engine exhaust pipe. 

Table 51 gives the sizes of barometric condensers as manufactured 
by prominent makers. 



TABLE 51. 
SIZE OF SIPHON CONDENSERS. 



Steam to be Condensed. 


Size Usually- 
Furnished, 
Inches. 


Steam to be Condensed. 


Size Usually 


Pounds per 
Hour. 


Pounds per 
Minute. 


Pounds per Hour. 


Pounds per 

Minute. 


Furnished, 
Inches. 


2,000 
3,000 
4,000 
5,000 
6,000 


33 
50 
66 

83 
100 


5 

7 
8 
9 
9 


8,000 
10,000 
15,000 
20,000 


133 
166 
250 
333 


10 
12 
14 
14 



Vacuum 26 inches; barometer 30 inches. 



410 



STEAM POWER PLANT ENGINEERING 



The diameter of the throat may be closely approximated by the 
empirical formula 

Diam. in inches = 0.0077 VWw, (88) 

in which 

W = weight of steam to be condensed per hour. 

w = weight of water required to condense one pound of steam. 

The maximum width of the annular opening for the admission of 
water may be obtained from the empirical formula 

Width in inches = Ww , (89) 

39,550 d 

in which 

d = diameter of the nozzle or bottom of 
the cone in inches. 
IF and was in (88) 

217. Ejector Condenser. — Fig. 206 shows 
a section through a Schutte exhaust steam 
"induction" condenser, illustrating the prin- 
ciples of the ejector condenser in which the 



EXHAUST 




Fig. 206. Schutte Ejector 
Condenser. 



Fig. 207. Piping for Schutte Ejector 
Condenser. 



momentum of flowing water ejects the discharge without the aid of the 
circulating pump. Exhaust steam enters the ejector through the open- 
ing marked " exhaust," passes through a series of inclined orifices and 



CONDENSERS 



411 



nozzles at considerable velocity, and, meeting the cooling water in the 
inner annular chamber, is condensed. The cooling water is drawn in 
continuously through the opening marked " water," by virtue of the 
vacuum formed, and sufficient velocity is imparted to the jet to dis- 
charge the combined mass of condensed steam, cooling water, and air 
against the pressure of the atmosphere. 

Adjustment for capacity is effected by raising or lowering the ram R 
by means of the wheel H. An adjustable sleeve controls the avail- 
able area of the exhaust inlet by covering more or less openings in the 
combining tube. When the cooling water is supplied under pressure 
the openings marked " steam " and are blanked. When water is 
taken under suction and water under pressure is available for starting, 
is blanked and opening marked 
"steam" is connected with the pres- 
sure supply. When water is taken 
under high suction and live steam is 
used for starting, inlet marked "steam" 
is connected to live steam and an over- 
flow check valve is placed at 0. Fig. 
207 gives an outline of the necessary 
piping for a condenser installation of 
this type. These condensers are made 
in all sizes conforming with exhaust 
pipe diameters of 1J to 20 inches. 
The same amount of cooling water is 
required as for jet condensing and 
vacua of 20 to 25 inches are readily 
obtained. 

Exhaust Steam Induction Condensers : 
Power, Dec, 1898, p. 14. Ejector Condenser: 
Hutton, Mechanical Engineering of Power 
Plants, p. Ill; Eng. News, Oct. 5, 1905, 
p. 360. 



218. Barometric Condensers.*— Fig. 

208 shows a section through a Weiss 
counter-current condenser, illustrating 
the principles of a barometric jet con- 
denser. The cooling water enters the 
upper part of the condensing chamber 
A through pipe iV and falls in cascades, 

as shown in the figure, to tail pipe B, from which it flows by gravity to 
the hot well. The exhaust steam enters chamber A through pipe D, 

* The author has been informed that the word " Barometric " in connection with 
jet condensers is the registered trade mark of the Alberger Condenser Company. 




Fig. 208. 



Weiss Counter-Current 
Condenser. 



412 



STEAM POWER PLANT ENGINEERING 



and, coming in contact with the cold-water spray, is condensed. The 
air is exhausted from the top of the condenser by a dry vacuum pump 
through pipe F. In flowing to the pump the air passes upwards 
through the water spray and its temperature is lowered to that of the 
injection water, thereby reducing the volume to be exhausted. Any 
moisture passing over with the air is separated at G before reaching 
the air pump, and flows out through the small barometric tube H. 
The cooling water is forced to the condenser chamber through pipe N 
by any positive displacement pump, the actual head pumped against 




Fig. 209. Section Through Condensing Chamber, Alberger Barometric Condenser. 



being the difference between the total height and that of a column of 
water corresponding to the degree of vacuum in the condenser. The 
main barometric tube or tail pipe B through which the water is dis- 
charged is 34 feet or more in length and is provided with a foot valve C. 
The counter-current principle permits a much higher temperature of 
hot well for the same degree of vacuum than does the parallel current, 
a hot- well temperature of 120 degrees and a vacuum of 27 inches being 
readily maintained. A small pipe K connecting the main condenser 



CONDENSERS 413 

with the small barometric tube H insures at all times a sufficient 
quantity of water in the small auxiliary hot well to seal the tube. 
The water from this auxiliary hot well flows over a weir, as indicated, 
into a counter-weighted bucket M, the latter having a hole in the bot- 
tom which allows the normal flow to escape. But in case a sudden 
heavy overload is thrown on the engines, and the adjustment is for a 
light load, the temperature of the discharge will reach the boiling point 
and an abnormal quantity of water will flow down the small barometric 
tube. This will cause the water to flow into the bucket much faster 
than the opening in the bottom can dispose of it; as a result the bucket 
will increase in weight and will open up a free-air valve L which 
reduces the vacuum two or three inches and raises the boiling point 
without " dropping " the vacuum entirely. E is the atmospheric relief 
valve. 

Fig. 209 shows a section through the condensing chamber of an 
Alberger barometric condenser. In principles of operation the con- 
denser is similar to the Weiss, but differs considerably in details. 
Exhaust steam enters at A and divides into two streams, one flowing 
directly to the inner chamber D, the other through the annular space E. 
Cooling water enters through B and is broken up into a fine spray 
by the serrated cone F, which is hung upon a long spring, thus auto- 
matically adjusting itself to the quantity of water entering the con- 
denser. After condensing the exhaust steam in the inner cylinder the 
partly heated spray of cooling water in falling is brought in contact 
with the exhaust steam which enters through the annular space. 
This process permits of a high hot-well temperature without affecting 
the degree of vacuum. The air which is not entrained by the cooling 
water and carried down the tail pipe collects under the spray cone F 
and ascends through the tubular support of the cone into the air 
cooler. This air cooler is simply a small chamber in which the non- 
condensable gases are cooled by a small portion of the circulating 
water before they are withdrawn by the air pump. The circulating 
water used for the purpose is forced into the cooling chamber 
through pipe K and falls through serrated openings in the bottom 
to the condenser proper. The air enters the chamber through 
these same openings, and is withdrawn by the air pump. Surround- 
ing the cooler is a separating space of large capacity to allow the 
subsidence of any entrained moisture before the air reaches the vacuum 
pump. 

Fig. 236 shows a typical installation of an Alberger condenser in 
connection with a cooling tower, and Fig. 226 that of a Weiss condenser 
in the Northwestern Elevated R. R. Power Station, Chicago. 



414 



STEAM POWER PLANT ENGINEERING 



Fig. 210 shows a section through the condensing chamber of a 
Worthington barometric condenser. The drawing is self-explicit. 



HAND WHEEL 



FROM.ENGINE 




Fig. 210. 



TO TAIL PIPE 

Section Through Condensing Chamber, Worthington Barometric Condenser. 



Fig. 211 shows a section through a Tomlinson barometric condenser. 
The air pump instead of discharging into the atmosphere is made to 
deliver into the tail pipe where the vacuum is still sufficient to support 
the column of water below the point of delivery. The effect produced 
is that of a two-stage air pump, the tail pipe becoming the second 
stage. Suitable by-pass valves enable the air pump to be discharged 
into the atmosphere or to be cut out entirely. (Power, February, 
1907, p. 94.) 

Fig. 211a shows the application of a centrifugal pump to the tail 
pipe of a barometric condenser. This permits of a very short tail pipe, 
as the pump takes the place of the barometric column. 

Counter-Current Condensers : Am. Elecn., Feb., 1905, p. 81; Power, March, 1905, 
p. 182, Jan., 1906, p. 44; Engr. U.S., Jan., 1906, p. 58; Hausbrand, . Evaporating 
and Condensing Apparatus, Chapter XX; Bulletin No. 6, Heisler Mfg. Co., St. 
Marys, Ohio. The Barometric Condenser : Power, Jan., 1907, p. 1. 



CONDENSERS 



415 




INJECTION 



Fig. 211. Tomlinson Type B Barometric Condenser. 



416 



STEAM POWER PLANT ENGINEERING 



As previously outlined, surface condensers may be divided into three 
general classes, (a) water cooled, (b) air cooled, and (c) evaporative. 

219. Water-Cooled Surface 
Condensers. — Water - cooled 
surface condensers are by far 
the most extensive in use and 
only occasionally are the con- 
ditions such as to warrant the 
installation of the other class. 
They are ordinarily classified 
as (1) single-flow, (2) double- 
flow, and (3) multi-flow. 

Fig. 212 shows a sectional 
elevation through a Baragwa- 
nath vertical condenser, illus- 




Fig. 211a. Centrifugal Pump Applied to the 
Tail Pipe of a Barometric Condenser. 



DISCHARGE 



STEAM 



trating the single- 
flow type. It con- 
sists essentially of 
a cast - iron shell 
provided with two 
heads, into which a 
number of one-inch exhaust 
brass tubes are ex- 
panded. Exhaust 
steam fills the shell 
and flows around 
and between the 
tubes, while the 
cooling water is 
caused to circulate 
through the tubes 
by means of a 
circulating pump. 
The steam is con- 
densed by contact 
with the tubes and 
drops to the bottom 
tube sheet, from 
which it is exhausted 
by the air pump. 
The circulating 
water flows through the tubes in one direction only, hence the name 
" single flow." To allow for the unequal expansion of shell and tubes 




TO AIR PUMP 



Fig. 212. Baragwanath Surface Condenser. 






CONDENSERS 



417 




418 



STEAM POWER PLANT ENGINEERING 



the two halves of the shell are provided with slightly thinner plates 
flanged outward, the flanges being bolted together with a spacing ring 
between them. This joint gives to the shell, in the direction of its 
length, a certain amount of elasticity which is sufficient to allow for the 
greatest possible elongation of the tubes without straining the tube ends 
and causing leakage. 

Fig. 213 shows a section through a Wheeler admiralty surface con- 
denser mounted on a combined air and circulating pump, illustrating 
the typical " double-flow " surface condenser. The condenser proper 
consists of a ribbed cast-iron chamber of rectangular section fitted 
with a number of small seamless drawn brass tubes through which 
the cooling water is forced by suitable means. The exhaust steam 
enters at the top and is prevented from impinging directly against the 
tubes by baffle plates, which serve also to distribute the steam more 
evenly over the cooling surface. The steam in passing between the 
tubes is condensed, and falls to the bottom of the chamber, from which 
it is removed, together with the entrained air, by a vacuum pump. 
The water chamber between the tube sheet and the head is divided into 
two compartments, as shown in the illustration, the partition being so 
arranged that the water flows first through the lower set of tubes and 
then through the upper set in the opposite direction. Thus the tem- 
perature of the cooling water increases as it rises, and reaches a maximum 
where the exhaust steam enters. Condensation begins as soon as the 
vapor enters the condenser, and the surfaces of the tubes are at once 
covered with a thin film of water flowing downwards from tube to 
tube. 




Fig. 214. Surface Condenser, C. H. Wheeler & Co. 



Fig. 214 gives the details of a C. H» Wheeler & Co.'s high- 
vacuum surface condenser. The condensing chamber is of the series- 
parallel type in which the water enters the top group of tubes, then 
passes to the middle section and finally through the bottom section- 



CONDENSERS 



419 



Connecting chambers are provided at the ends of the shell as illustrated. 
This construction of water chamber keeps the condenser completely 
filled with cooling water at all times. The inlet is at the bottom but the 
water is carried up through the annular chamber to the top of the tubes. 




Fig. 215. Weighton Multi-Flow Surface Condenser. 



Fig. 215 shows a section through a multi-flow surface condenser 
designed by Professor R. L. Weighton. The condenser has three 
compartments separated by two diaphragms inclined to the hori- 
zontal. Each compartment is fitted with a number of brass tubes 
three-fourths inch in diameter by four feet in length, spaced one 
and one-eighth inches between centers. The cooling water circulates 
through the tubes five times, giving an effective length of 20 feet. The 
notable features of the condenser are abolishment of steam space, and 
compartment drainage of condensed steam. Mere passages of such shape 



420 STEAM POWER PLANT ENGINEERING 

and section as will insure distribution of the steam over the entire surface 
are used instead of the large steam space usually associated with surface 
condensers. Each compartment is separately drained to the air pump, 
so that the surfaces in the lower compartments are unimpeded in their 
condensing action by the condensed steam from the upper compartments 
flowing over them. Referring to Fig. 215: Exhaust steam enters the 
condenser at A and flows toward the hot well H. The greater part 
of the steam is condensed in the first section of the condenser, and the 
condensation is drained directly to the hot well. The balance of the 
condensation takes place in the remaining sections, the condensed steam 
being withdrawn from each section. The wet-air pump withdraws the 
condensed steam and non-condensable gases through opening P. Cool- 
ing water enters at / and leaves at 0. An exhaustive series of tests on 
a condenser of this type credit it with a much higher efficiency than 
the ordinary single or double-flow apparatus. (See " The Efficiency of 
Surface Condensers," Proc. Institute of Naval Architects, March, 1906; 
also Engineer, London, April 27, 1906.) 

220. Cooling Water, Surface Condensers. — The amount of cooling 
water required per pound of steam in a surface condenser is dependent 
upon the vacuum, the temperature of the condensed steam, and the 
range in temperature of the cooling water; it may be closely approxi- 
mated from the formula 

W- A ~ Tl + 32 . (90) 

where 

\ = total heat of the exhaust steam above 32 degrees F. 
T x = temperature of the condensed steam. 
T = temperature of the injection water. 
T 2 = temperature of the discharge water. 

W= pounds of injection water necessary to condense one pound 
of steam. 

Example : Required the quantity of cooling water necessary to con- 
dense one pound of steam under the following conditions: Initial tem- 
perature of the cooling water 60 degrees F.; final temperature 100 
degrees F.; vacuum 26 inches, referred to 30-inch barometer. Here 
X = 1120 B.T.U., T = 60, T 2 = 100. 

= 1120-110 + 32 _ 
100 - 60 

That is, the ratio of cooling water to condensed steam is 26.0 to 1. 
In turbine practice where vacua as high as one-half pound absolute are 



CONDENSERS 



421 



obtained, the ratio of cooling water to condensed steam is nearly twice 
this quantity. For example, if a vacuum of 28.92 inches is desired 
with the barometer at 29.92 and the range of the circulating water tem- 
perature is 70 to 50 degrees and the temperature of the hot well 80 
degrees, the ratio will be 

1106-80 + 32 



W*= 



70-50 



= 52.9. 



In determining the amount of cooling water it is well to bear in 
mind that in the ordinary condenser of the single or double-flow type 



150 

140 

130 

120 

110 

100 

90 

80 

TO 

60 

50 

40 



















A New Type with Cores and Spray and 

Dry Air Pump 
B New Type without Cores and Spray 

Ordinary Pump 




































C Old Type, Ordinary Pump 






















\A 






























B^ 










+3 


















<^.C 














!3 

ft 
































3 
Pi 






























\ 


S3 

\ 








Relation between Hot-Well Temperature 
and Vacuum in Surface Condensers 






\ 


\ 
\ 


































\ 
\ 
































\ 



23 24 25 .26 27 

Vacuum Referred to 30 Inch Barometer 
Fig. 216. 



the temperature of the condensed steam will be from 10 to 20 degrees 
lower than that corresponding to the degree of vacuum in the con- 
denser, and that the temperature of the condensing water at the dis- 
charge point will be from 5 to 10 degrees lower than the temperature 
due to the vacuum. 

With well-designed condensers of the multi-flow type the temper- 
ature of the hot well may be from 3 to 5 degrees higher than the tem- 
perature due to the vacuum, and the temperature of the condensing 
water at the discharge point may be equal to or slightly higher than 
that due to the vacuum. (Proc. Inst, of Naval Arch., March, 1906.) 
(See Fig. 216.) 

221. Extent of Water-Cooling Surface. — Theoretically the opera- 
tion of a surface condenser is divided into two periods, (1) the period 



422 STEAM POWER PLANT ENGINEERING 

of condensation during which the heat of vaporization at the observed 
pressure is removed and (2) the period of cooling during which the 
temperature of the condensed steam is reduced. In order to determine 
accurately the extent of cooling surface it would be necessary to cal- 
culate the heat transmission for each of the two periods. In practice, 
however, it is assumed that condensation and cooling take place simul- 
taneously, and that the mean temperature difference is a direct function 
of the temperature corresponding to the exhaust steam in the con- 
denser and that of the condensed steam and cooling water. The error 
in these assumptions has only a slight influence on the estimation of the 
cooling surface and is entirely lost sight of in the liberal factor allowed 
in practice. 

Let & = cooling surface in square feet. 

\ = total heat above 32 degrees of the exhaust steam at con- 
denser pressure. 

T = initial temperature of the circulating water. 

T 2 = final temperature of the circulating water. 

T x = final temperature of the condensed steam. 

Ti = temperature of the exhaust steam at condenser pressure. 

U = coefficient of heat transmission, B.T.U. per hour, per degree 
difference in temperature, per square foot of cooling sur- 
face. 
d = mean difference in temperature between T s and T 2> and T . 

W = weight of condensed steam per hour. 

d = T *~ T ° "(see equation (118), Chap. XII); 

and since the heat absorbed by the cooling water is equal to the heat 
given up by the steam, 

SUd = W{X-(T 1 -32)}. (91) 

s =W (X-T 1+ 32) . 
Ud 

Whitham (" Steam-Engine Design,' ' p. 283) uses the arithmetic mean 

T + T 
d = T s °— - — 2 instead of the mean as determined from (118). 

Equation 118 is based on the assumption that the fluid on each side 
of the tube is homogeneous, which is far from being true in the case of 
the air-steam mixture in a condenser, and for this reason many designers 
prefer to use the simpler arithmetic formula. 

The coefficient of heat transfer, U, as used in above equations refers 



CONDENSERS 



423 



tc the mean or average values for the entire surface since the actual heat 
transmission varies widely for different parts of the condenser; thus 
the actual value of U varies from over 1000 in the first few rows of the 
tubes (where the steam comes directly into contact with the cooling 
surface) to less than 50 in the bottom row (where the tubes are practi- 




Fig. 216a. Application of Weighton Dry-tube Surface Condenser to 
Vertical Marine Engine. 

cally submerged in water of condensation) and to 3 or less for the tubes 
surrounded only by air. 

Prof. Josse of the Royal Technical School, Charlottenburg, after an 
exhaustive investigation of the subject found that the actual value of 
U varied with 

(1) The material, thickness, shape and cleanliness of the tubes. 

(2) The velocity of the water through the tubes. 



424 



STEAM POWER PLANT ENGINEERING 



(3) The velocity of the steam against the tubes. 

(4) The percentage of air in the steam surrounding the tubes. 

(5) The extent of submersion of the steam side of the tubes. 

Some of the results of his investigations are shown in Figs. 216, b, 

c, and d. See also Power and 
' ~ Engr., Feb. 2, 1909. 

The effect of thickness, ma- 
terial, etc., of condenser tubes is 
so small in the ultimate result 
and the choice and arrangement 
are so largely determined by 
practical consideration that they 
may be neglected. 

The value of U increases ap- 
proximately as the square root 
of the velocity of the water flow- 
ing within the tube, so that in- 
crease in water velocity effects a 
substantial increase in the heat 
transmission ; but the resistance 
encountered by the circulating 
water increases as the square of 
the velocity, and the power con- 
sumed in pumping the water in- 
creases as the third power of the 
velocity, so that a point is soon 
reached where the gain on the 
one hand may be offset by the 
loss on the other. 

A study of a number of instal- 
lations gave 




12 3 4 5 6 7 

Rate of Flow of Cooling Water- Feet per Second 

Fig. 216b. 



Old-style surface condenser, 

V = 30 to 240 ft. per min., average 90. 
Modern dry tube surface condenser, 

V = 120 to 360 ft. per min., average 240. 

From the curves in Figs. 126b and 126c it will be seen that air is an 
excellent heat-insulating material; hence, the greater the amount of 
air entrained with the steam the lower will be the coefficient of heat 
transmission. The necessity of removing the air as fast as it accumu- 
lates is at once apparent. 

In the older types of surface condensers the water of condensation 
from the upper rows of tubes is permitted to fall on the rows immedi- 



CONDENSERS 



425 



Heat Transference for Air 




10 20 30 40 50 60 70 
Mean Rate of Flow of Air —Feet per Second 

Fig. 216c. 



ately below, the water increasing in volume as it passes the successive 

banks of tubes until it completely envelops them. The coefficient U 

varies from 1000 or more in the 

upper row to less than 50 in the 

lower, giving a mean value of 

approximately 250 to 350 for the 

entire surface. In estimating 

the extent of cooling surface for 

a condenser of this type an 

average figure for plain brass 

tubes with water velocities of 50 

to 100 feet per minute is U = 250. 

For a velocity of 100 to 240 feet 

per minute U may be taken 50 per 

cent greater than these figures. 

When the tubes are clean a much 

higher value may be taken, but 

a liberal factor is usually allowed 

for possible variation in the con- 
dition of operation. 

In the modern dry-tube surface condenser, designed along the lines 

of the one described in paragraph 221, in which the water of condensa- 
Heat Transference for Air tion is withdrawn as rapidly as 

it is formed, mean values of 
U = 800 to 900 are not unusual. 
In estimating the extent of 
cooling surface for condensers 
of this type an average value 
of U is 600 with water veloci- 
ties of 4 to 5 feet per second. 

Example: Standard Type of 
Surface Condenser : — Required 
the number of square feet of 
cooling surface per I.H.P. neces- 
sary to condense the steam from 
an engine operating under the 
following conditions : Engine 

uses 20 pounds of steam per I.H.P. -hour, vacuum 26 inches with 

barometer at 30; temperature of cooling water at 60 degrees. 
Here X = 1115 and T s = 126 (from steam tables), 

T = 60, 

T 1 = T s - 10 = 116. 



w Q 

go t 

K £4 

K a* 
A 













,> 


<* 










e<py / ' 
























A 


#•> 










o) 


V 


f^ 


<g2> 








*f 


















5.25 **• 


perSecN 





















25 20 15 10 5 

Air Pressure "-Inches of Mercury 

Fig. 216d. 



426 STEAM POWER PLANT ENGINEERING 

In this type of condenser average practice gives a temperature 
difference of approximately 10 degrees between the temperature of the 
hot well and that corresponding to the degree of vacuum. 

T 2 = T s - 15 - 101. 
Any value may be fixed upon for T 2 greater than T and less than T s . 
The nearer T 2 is to T the greater must be the quantity of circulating 
water per unit of time for a given rate of condensation. On the other 
hand, the nearer T 2 is to T s the less is the mean temperature difference 
d and hence the greater must be the cooling surface for a given coeffi- 
cient of heat transmission. When water is cheap and the head pumped 
against is small T 2 should be given a lower value than when water is 
costly and the discharge head is large. Average engine practice, with 
conditions as stated, gives T 2 a value of approximately 15 degrees less 
than that corresponding to the degree of vacuum. 
The logarithmic mean is, equation (118), 

101 - 60 
*- 126 - 16 = 42A 

ge 126 - 101 
The arithmetic mean gives 

, ™ 60 + 101 ._ 
a = 126 » = 45.5. 

Substitute the value of d in equation (92) and assume U = 250, the 
figure commonly used for this type of condenser. 
_ _ 20(1115- 116 + 32) _ 
* " 250 X 42.4 ~~ iy4 ' 

or say two square feet per I.H.P. of engine. 

Surface condensers of this type are ordinarily rated on a basis of two 
square feet per I.H.P. 

Example: Dry-tube Multi-flow Surface Condenser: — Required the 
number of square feet of cooling surface per kilowatt necessary to 
condense the steam from a steam turbine operating under the following 
conditions: Turbine uses 15 pounds of steam per kw.-hour; vacuum 
28.5 inches, referred to 30-inch barometer; temperature of cooling 
water 70 degrees. 

Here i = .9X 1100 = 990. 

The total heat of dry steam corresponding to an absolute pressure 
of 1.5 inches is 1100, but in the case of high vacuum turbine practice 
the steam entering the condenser is far from being dry, the quality 
varying from 0.80 to 0.95, depending upon the quality of the steam at 
admission. An average figure is 0.9. 

T s = 92, T = 70, T t = T s - 4 = 88. 






CONDENSERS 



427 



In this type of condenser the hot -well temperature varies from 
T x = T s to T x = T s - 8. 



T 

± s 



In this type 7\ 



varies from T 2 = T s 



= 87. 
to 7\, = 



10. 



87 - 70 



log 



90 - 70 

92 - 87 



11.5. 



= 2.02, 



70 -I- 87 
Arithmetic mean gives d = 92 ■= — = 13.5. 

Substitute the value of d in equation (92) and assume U = 600, the 
figure commonly used for this type of condenser. 
_ 15 (990 - 88 + 32) 
600X11.5 

or say 2 square feet per kw. of generator. There is no standard 
rating of surface condenser for steam turbine work because of the wide 
variation in operating conditions. A study of a number of modern 
installations gives 

1 to 2.5 sq. ft. per kw. for large turbo-generators using 

dry tube surface condensers. 
2.5 to 4 sq. ft. per kw. for small turbo-generators using 
standard surface condensers. 
Professor Weighton found from his experiments that a surface con- 
denser constructed on the lines of the one described in paragraph 221 in 
conjunction with dry-air pumps, was capable of condensing 20 pounds 
of steam per square foot of surface per hour and maintained a vacuum 
of 28^ inches (referred to a 30-inch barometer), and this with a cooling- 
water consumption of 24 pounds per pound of condensed steam; with 
an inlet temperature of 50 degrees F. a condensation of 35 pounds of 
steam per hour per square foot of cooling surface was effected at a 
ratio of 28 pounds of cooling water per pound of steam, the vacuum 
remaining 28^ inches. See Fig. 216. (Engineering Record, May 19, 1906, 
p. 615.) 

EXAMPLES OF MODERN CONDENSER PROPORTIONS. 



Name of Station. 



Commonwealth Edison Co.: 

Northwest Station 

Quarry Street 

Fisk Street 

*59th St., Interborough, N. Y 

Metropolitan St. Ry., Kansas City 



Size of Turbo- 
Generators. 



20,000 
14,000 
12.000 
15.000 
10.000 



Sq. Ft. of 

Condenser 
Surface. 



32,000 
25,000 
25,000 
25.000 
22,000 



Sq. Ft. of 
Surface 
per Kw. 



1.60 
1.79 
2.08 
1.67 
2.20 



* Combined Engine and Low-pressure Turbine. 

Surface Condenser Air Pumps. — See paragraphs 284-291. 



428 



STEAM POWER PLANT ENGINEERING 



TABLE 52. 

SQUARE FEET OF COOLING SURFACE NECESSARY TO CONDENSE AND COOL 
ONE POUND OF STEAM PER MINUTE. 

(Barometer 29.92.) 



Temp. 


Vacuum 24". 
Temperature of Steam 141°. 


Temp, 
of In- 
jection 
Water. 


Vacuum 25". 
Temperature of Steam 134 





of In- 
jection 
Water. 


Temperature of Hot Well. 


Temperature of Hot Well. 




110 


115 


120 


125 


130 


105 


110 


115 


120 


125 


40 
50 
60 
70 
80 
90 


3.1 
3.3 
3.7 
3.9 
4.1 
4.4 


3.3 
3.5 
3.8 

4.1 
4.4 

4.8 


3.5 
3.7 
4.0 
4.3 
4.7 
5.1 


3.7 
4.0 
4.2 
4.6 
5.0 
5.5 


4.0 
4.3 
4.6 
5.0 
5.5 
6.0 


40 
50 
60 
70 
80 
90 


3.3 
3.5 
3.8 
4.1 
4.5 
5.1 


3.5 

3.7 
4.0 

4.4 
4.8 
5.4 


3.7 
4.0 
4.3 
4.7 
5.2 
5.8 


4.0 
4.3 
4.6 

5.1 
5.6 
6.2 


4.4 
4.7 
5.1 
5.5 
6.1 
6.9 


Temp. 


Vacuum 26". 
Temperature of Steam 125°. 


Temp. 

of In- 
jection 
Water. 


Vacuum 27". 
Temperature of Steam 114°. 


of In- 
jection 
Water. 


Temperature of Hot Well. 


Temperature of Hot Well. 




100 


105 


110 


115 




90 


95 


100 


105 




40 
50 
60 

70 
80 


3.6 
3.8 
4.2 
4.6 
5.1 


3.9 
4.2 
4.6 

5.1 
5.6 


4.2 
4.6 
5.0 
5.4 
6.1 


4.6 
5.0 
5.4 
6.0 
6.7 




40 
50 
60 
70 

80 


4.1 
4.4 

4.8 
5.4 


4.4 
4.7 
5.2 
5.8 
6.7 


4.7 
5.1 
5.G 
6.3 
7.2 


5.1 
5.6 
6.2 
7.0 
8.0 





Temp. 


Vacuum 28". 
Temperature of Steam 100°. 


Temp. 

of In- 
jection 
Water. 


Vacuum 29". 
Temperature of Steam 77 





of In- 
jection 
Water. 


Temperature of Hot Well. 


Temperature of Hot Well. 




75 


80 


85 


90 




60 


65 


70 






40 


4.6 
5.0 
5.5 


4.9 
5.4 
6.1 


5.3 
5.8 
6.6 

7.7 


5.8 
6.4 
7.3 
8.5 




35 
40 
45 
50 


6.3 
6.8 
7.4 


6.9 
7.4 
8.0 

8.9 


7.7 
8.3 
9.0 
9.9 






50 






60 

70 





















Final temperature of injection water assumed to be 10 degrees lower than that of the hot-well. 

222. Dry-Air Surface Condensers (Forced Circulation). — Where 
water is very scarce and the feed supply is reclaimed by condensing the 
exhaust steam, water-cooled condensers may be prohibitive in cost of 
operation, even when combined with cooling tower or other water-cool- 






CONDENSERS 



429 



ing device, since the latter involves a loss of water approximately 
equivalent to the amount of steam condensed, due to evaporation. 

Under these conditions air cooling has been successfully adopted. 
In the city of Kalgoorlie, West Australia, an electric station of 2000- 
horse-power capacity is equipped with air-cooled surface condensers. 
The condensers have been in use five years (1906), and have given 
excellent service with very little expense and maintenance. The con- 
denser consists of a large number of narrow chambers constructed of 
thin corrugated sheet-steel plates spaced J inch between centers. Each 
chamber has 1345 square inches of cooling surface. Fifty-one of 
these chambers are grouped into a compartment and 15 compartments 
constitute a section. Each section is equipped with three motor-driven 
fans 7 feet in diameter and running normally at 320 r.p.m. In all 
there are six sections, giving a total cooling surface of 45,000 square 
feet. The steam consumption of the main engines is 16 to 16.5 pounds 
per I.H.P. hour at rated load. At full load the fans require 130 kilo- 
watts, or approximately 10 per cent of the station output. The average 
vacuum obtained is about 18 inches throughout the year and ranges 
from inches on very hot days to 22 inches in cooler weather. The 
following figures, based on actual observation, show the effect of tem- 
perature of the external air on the vacuum when condensing 32,000 
pounds of steam per hour (the rated capacity of the condenser) . 



Temperature Ex- 


Vacuum, Inches 


Temperature Ex- 


Vacuum, Inches 


ternal Air, 


(referred to 30-Inch 


ternal Air, 


(referred to 30-Inch 


Degrees F. 


Barometer) . 


Degrees F. 


Barometer) . 


42.8 


22 


96.8 


9.6 


50 


21.2 


100.4 


7.6 


60.8 


20 


107.6 


3.6 


68 


18.4 


113 





78.8 


16 







Air-Cooled Surface Condensers : Engineering News, Oct., 1902, p. 271 ; ibid., Vol. 
49, p. 203. 

223. Quantity of Air for Cooling (Dry-Air Condenser). — The volume 
of air, under atmospheric conditions, necessary to condense steam to 
any given temperature may be determined as follows: 

Let X = total heat, above 32 degrees F. of the steam at condenser 
pressure. 
T s = temperature of the vapor in the condenser. 
T x = temperature of the condensed steam. 
t = temperature of the air entering condenser. 
t x — temperature of the air leaving condenser. 
V = volume of air in cubic feet necessary to condense and cool 
one pound of steam. 



480 STEAM POWER PLANT ENGINEERING 

B = specific weight of air under atmospheric conditions. 
C = specific heat of air under atmospheric conditions. 
d = mean temperature difference between the air and steam. 
S = cooling surface in square feet. 

U = coefficient of heat transmission, B.T.U. per square foot per 
degree difference in temperature per hour. 

Since the heat absorbed by the air must be equal to the heat given 
up by the steam, neglecting radiation we have 

VBC (t.-t) = \-T\ + 32, (93) 

from which 

V = ^,+32 . (94) 

For practical purposes C may be taken as the specific heat of dry air, 
the error due to this assumption being negligible even if the air is 
saturated with moisture. 

Example : How many cubic feet of air are necessary to condense and 
cool one pound of steam under the following conditions : Vacuum 20 
inches; temperature of entering air, leaving air, and condensed steam, 
60, 110, and 140 degrees F. respectively? 

Here A = 1131 (from steam tables). 

t x = 110, T 1 = 140, t = 60, C = 0.2377, B = 0.075. 

Substituting these values in equation (94), 

jr 1131-140 + 32 11W1 u- * + * 

1 = 0.075X0.2377(110-60) = 115 ° ° UblC feet ° f air neCGSSary t0 
condense one pound of steam under the given conditions. 

The proper area of cooling surface depends upon the value of the 
coefficient of heat transmission, which varies with the velocity and 
humidity of the air and character of the cooling surface. Accurate 
data are not available on this point. 

A few experiments made at the Armour Institute of Technology 
gave values of U = 10 to 25 B.T.U. per hour, per square foot, per 
degree difference in temperature for air velocities of 500 to 4000 feet 
per minute for corrugated steel sheeting J inch thick. Hence, sub- 
stituting in equations (94) and (92) we get, for the above example,. 
S = 1.5 square feet of cooling surface per pound of steam condensed 
per hour for air velocity of 4000 feet per minute, and S = 3.7 square 
feet for a velocity of 500 feet per minute. 

224. Saturated-Air Surface Condensers (Natural Draft). — Fig. 217 
shows vertical and horizontal sections of a Pennel saturated-air surface 



CONDENSERS 



431 



condenser. The apparatus consists of an upright cylindrical shell 
containing a number of vertical 4-inch steel tubes through which air is 
drawn by natural draft. A centrifugal pump circulates about one- 
half gallon of water per horse power per minute from a cistern below the 
condenser. The water flowing over the upper tube sheet and then 
descending the tubes by gravity forms a film over their entire interior 
surface. 




Horizontal Section. Section on AB. 

Fig. 217. Pennel Saturated- Air Surface Condenser. 



The condensing action is as follows: The current of exhaust steam 
entering the side of the shell at A is caused by suitable baffle plates to 
circulate among the tubes, and in condensing gives up its latent heat 
to the water film, which wholly or partially evaporates, saturating the 
ascending current of air at its own temperature. The upward current 
of hot vapor-laden air carries off the heat into the atmosphere. The 
cooling water which is not evaporated and lost to the atmosphere falls 
into the cistern below to be again taken up by the circulating pump, 
the water level in the cistern being kept constant by a float governing 
a valve on the supply pipe. The non-condensable gases collect at C, 
where they are removed by the dry -air pump, while the condensed steam 
is drawn off from the bottom tube sheet by the vacuum pump and 
discharged into the hot well. An excellent feature of this device is 
that the film of water on the cooling surface is secured without inter- 
ference with the ascending air currents and also without the use of 
sprays through small orifices likely to become clogged with rust or 
sediment. Where the recovery of the condensed steam is essential and 
a high vacuum of secondary importance, condensers of this type have 
proved to be good investments on account of the low first cost. 



432 



STEAM POWER PLANT ENGINEERING 



Table 53 gives the results of a test of a condenser of this type, taking 
steam from a 30 x 58 x 48 engine running at 45 r.p.m. (Power, December, 
1903, p. 672; West Elect, May 19, 1900, p. 323.) 



TABLE 53. 

TEST OF PENNEL SATURATED-AIR SURFACE CONDENSER. 

Duration of trial .9 hours 

Average steam pressure at engine by gauge 139.8 pounds 

Average vacuum, mercury column 17.5 inches 

Average temperature in condenser 123.7 degrees F. 

Average temperature of circulating water 116.4 do 

Average temperature of city water 52 do 

Average temperature of outside air . . 62 do 

Average temperature of saturated air 106 do 

Average draft in stack of condenser 1.1 inches 

Average humidity of outside air 67 per cent 

Average amount of steam condensed per hour 7950 pounds 

Average amount of circulating water used per hour 114,660 pounds 

Average amount of city water used per hour 3462 pounds 

Pounds of city water per pound of steam 2.3 

Pounds of circulating water per pound of steam 14.4 

Average horse power of engine 569.7 

Steam, pounds per I.H.P. per hour 13.95 

Horse power required to run air pumps 10 . 5 

Horse power required to run circulating pumps 3.0 

Condensing surface, square feet 3900 

Pounds of steam condensed per square foot surface per hour 2038 

Barometer 28.58 inches 

Vapor tension corresponding to 123.7 degrees 3.82 inches 

Per cent of main engine steam used by auxiliaries 2.38 

218 illustrates the Pennel 



a 



b 



Fig. 
" flask " type of atmospheric 
condenser. The exhaust steam 
enters below and follows the zig- 
zag course bounded by the inter- 
nal stay channels, condensing as 
it goes and driving before it the 
non-condensable gases to the out- 
let at the top. The condensed 
steam gravitates to the bottom 
and thence to the hot well. The 
I 4^4" V v> top of the flask is trough shaped 
" and causes the cooling water to 

Fig. 218. Pennel Flask Type of Saturated- fl()W down the gides of the flagk 
Air Surface Condenser. . 

in a thin stream. The portion 
of the cooling water not evaporated collects at the bottom of the flask 
and flows to the cooling-water reservoir. 



3E 



CONDENSERS 433 

225. Evaporative Surface Condensers. — An evaporative surface con- 
denser consists of a number of copper, brass, wrought- or cast-iron 
tubes arranged horizontally or vertically and connected to manifolds 
or chambers at each end. The exhaust steam passes through the 
tubes and a thin film of water is allowed to flow over the external 
surfaces. The cooling effect is brought about by the evaporation of 
part of the circulating water, and the general principle of operation 
is the same as that of the saturated-air condenser described above. 
Evaporation is sometimes hastened by constructing a flue over the 
tubes, thereby creating a natural draft, or by means of fans. With 
horizontal cast-iron tubes and natural draft, vacua from 23 to 27 inches 
are readily maintained with a cooling surface of approximately eight- 
tenths square foot per pound of steam condensed per hour. With 
vertical brass tubes and fan draft 8 pounds of steam per hour per 
square foot of cooling surface is not an unusual figure. The amount 
of cooling water evaporated per pound of steam varies from eight- 
tenths to one pound, depending upon the draft. The power necessary to 
operate the pumps and fans varies from 1 to 4 per cent of the total 
output of the plant. For an interesting discussion of evaporative 
condensers the reader is referred to the admirable article by Oldham 
in the Proceedings of the Institute of Mechanical Engineers, 1899, and 
reproduced as a serial in Engineering (London), April 28 to June 30, 
1899. The following test of a vertical cast iron tube evaporative 
surface condenser (Table 54) will give some idea of the performance 
of this type of condenser. This condenser consisted of two rows of 
4-inch vertical cast-iron pipes connected at the top by U bends and at 
the bottom by cast-iron manifolds. A perforated iron trough dis- 
tributes the water over the center of the bend and causes it to flow in 
a thin stream over the surface of the tubes. A wet-air pump is used 
for withdrawing the condensed steam and air. No fan is used for 
hastening evaporation.* 

Evaporative Condensers: Engr.,Lond.,May 5, 1899, pp. 432, 442, 447 ; Engineering, 
May 19, 1899, p. 661, June 2, 1899, p. 721, June 30, 1899, p. 861; Trans. A.S.M.E. 
14-696; Power, Sept., 1904, p. 542; Prac. Engr. U.S., June, 1910, p. 346. 

226. Location and Arrangement of Condensers. — In the modern 
power house one sees two general arrangements of condensers and 
auxiliaries: 

1. The independent or subdivided system/in which each engine or 
turbine is provided with its own condenser, air and circulating pumps. 

2. The central system, in which the condensers and auxiliaries are 
grouped together. Ordinarily one condenser suffices for all engines. 

* See end of paragraph 236 for evaporated surface condenser calculations. 



434 



STEAM POWER PLANT ENGINEERING 



TABLE 54. 

TEST OF A CAST-IRON, VERTICAL-TUBE, EVAPORATIVE SURFACE CONDENSER, 

NATURAL DRAFT. 



Date 

Weather 

Barometer 

Temperature of air 

Cooling surface, external 

Duration of trial, minutes 

Weight of steam condensed, pounds 

Boiler pressure 

Weight of water in circulation 

Weight of fresh water added 

Vacuum in condenser 

Initial temperature of circulating water 

Final temperature of circulating water 

Temperature of " make up " water 

Temperature of water in hot well 

Weight of steam condensed per hour, pounds . . . 

Weight of water circulated per hour, pounds 

Weight of " make-up " water added per hour. . . 
Weight of steam condensed per square foot of 

cooling surface per hour 

Weight of "make-up" water per pound of steam 

condensed, pounds 



Sept. 12 


Sept. 13 


Wet 


Fine 


29.8 


29.5 


? 


60 


272 


272 


99 


115 


800 


800 


60 


60 


1830 


1830 


600 


640 


23.36 


24.1 


117.5 


113.9 


128.4 


125 


58 


58 


136.5 


131.8 


485 


427 


6786 


? 


364 


334 


1.8 


1.54 


0.75 


0.80 



226a. The Independent System. — The condenser is usually placed 
close to and below the engine so that all condensation may gravitate 
into it. Figs. 219 and 221 show an application of this system with 
jet condensers. Here each condenser receives its supply of cooling 
water from a main injection pipe and discharges into a main overflow 
pipe. The exhaust pipe leading to the condenser is by-passed through 
a suitable atmospheric relief valve to a main free exhaust header so 
that the engine may operate non-condensing in case the vacuum breaks 
or the condenser is cut out. The chief feature of this arrangement is 
its flexibility, as each unit is complete in itself and independent of 
the others. By far the greater number of central stations are equipped 
with independent condensers. 

Occasionally a jet condenser is located on the same level with the 
engine or even above it, Fig. 222, but such a location should be avoided 
if possible, as it usually necessitates a larger number of bends and 
joints in the exhaust pipes than the basement arrangement, and 
increases the possibility of air leakage. If the exhaust pipe does not 
drain directly into the condenser, the lowest point in the piping should 
always be provided with a drip which should be opened when the 
engine is shut down, as condensation and leakage are apt to fill the 
pipe with water if the engine stands for any length of time. The end 



CONDENSERS 



435 







DISCHARGE -*- 



ATMOSPHERIC RELIEF VALVE 



SJijigSjUttM^ ' 



Fig. 219. Jet Condenser located below Engine-Room Floor. 




Fig. 220. Surface Condenser located below Engine-Room Floor. 



436 




Fig. 221. Surface Condenser, Installed in the Suction Line of a Pumping Engine. 




n — i r^-r 

Fig. 222. Jet Condenser located above Engine-Room Floor. 



Fig 223 Typical Arrangement, Westinghouse-Leblanc 
Condenser and Curtis Turbine 



-ALL FOREIGN SOB STAN 
CES SUCH AS LEAVES, 

«&TIC«S STRAW.ETC. 



TSfcSJTSfe 




CONDENSERS 



487 




& 

6 

to 

a 

I 



438 



STEAM POWER PLANT ENGINEERING 



of the drip should be connected so that water cannot be drawn back 
through the drip pipe and into the engine cylinder. The length of 
exhaust pipe and particularly the number of bends between engine and 
condenser should be kept as small as possible, otherwise the engine 
may not derive the full benefit of the vacuum in the condenser. A 
case is recorded where the exhaust piping and appurtenances in con- 
nection with a 5000-horse-power engine caused a drop of several inches 
in vacuum between condenser and exhaust opening of the low-pressure 
cylinder. {National Engineer, December, 1906, p. 10.) The wet-air 
pump must always be located below the condenser chamber so that the 
condensation may gravitate to it. 




Fig. 225. Plan of Piping for Engine and Condenser, Des Moines City Ry. Co. 

Fig. 252 shows a surface condenser installation in connection with a 
vacuum or primary heater. 

Fig. 236 shows an application of a barometric condenser to a 
vertical engine installation. 

Fig. 220 shows the arrangement of a surface condenser with com- 
bined air and circulating pump in connection with a horizontal cross 






CONDENSERS 439 

compound engine. The condenser and appurtenances are placed below 
the engine, thereby permitting the condenser to be closely connected 
to the engine. 

Fig. 221 shows the arrangement of a surface condenser in connec- 
tion with a pumping engine. The condenser is placed in series with 
the pump suction. 

227. Central Systems. — In the central condensing systems the con- 
denser is located at any convenient point and the exhaust from all the 
engines piped to it. Any arrangement of condenser and auxiliary 
machinery may be adopted which will favor the lowest cost of installa- 
tion and expense of operation. Except where continuity of opera- 
tion is absolutely essential, only one circulating pump and one air pump 
are installed. This reduces the number of auxiliary pumps and appliances 
to a minimum, with a consequent decrease in first cost and maintenance. 
With properly designed exhaust piping the condenser may be located 
at a considerable distance from the engine without undue loss of vacuum. 
At the Cambria Steel Works, Johnstown, Pa., the maximum drop 
between condenser and engine is only three-quarters of an inch and the 
distance between them is about 1000 feet. 

Central condensers have found great favor in power plants in which 
the individual units are subjected to extreme variations in load, as in 
rolling mills. At the works of the Illinois Steel Company, South Chicago, 
111., one condenser takes care of the steam from 15,000 horse power of 
engines in the rail mill, and another condenses the steam from the 
15,000 horse power of engines in the Bessemer steel mill. A notable 
installation of this system in connection with street-railway work is 
in the power house of the Northwestern Elevated Company, Chicago, 
where a single condenser takes care of the exhaust steam of five engines, 
11,000 horse power in all. Fig. 226 shows the general arrangement of 
this installation. 

For a comparison of the advantages and disadvantages of the inde- 
pendent and central systems see Engineering Magazine, October, 1900, 
p. 56, Engineering, London, June 23, 1899, p. 615, and Engineering, July 
17, 1903. 

Centralization of Steam-Condensing Plant : Eng. Mag., Oct., 1900, p. 56 ; Iron Age, 
Jan. 7, 1904; Revue Technique, Feb. 25, 1903. 

Five Thousand H.P. Surface-Condensing Plant : Engr., Lond., May 23, 1903. 

Aurora & Elgin R.R. Condenser Plant : Engr. Rec, Vol. 47, p. 153. 

Condensing Apparatus of Manhattan Elevated Power Plant: Power, Aug., 1903, 
p. 411. 

Interborough R.R. Condenser Plant : St. Ry. Jour., Oct. 8, 1904. 

New York Rapid Transit Condenser Plant : Power, June, 1903, p. 283. 

Worthington Surface Condensers for Metropolitan Power Station : Power, June, 
1901, p. 15. 



440 



STEAM POWER PLANT ENGINEERING 



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CONDENSERS 



441 



228. High-vacuum Systems. — The average reciprocating engine 
gives its best commercial economy at a vacuum of approximately 26 
inches (referred to a 30" barometer), and the ordinary standard jet or 
surface condenser has been designed to meet this requirement. At 
the time of the introduction of the steam turbine it was discovered that 
a very high vacuum would improve turbine economies to an extent 
hitherto impossible when applied to reciprocating engines. This con- 
dition naturally created an era of development among the condenser 




Fig. 226a. Condenser Installation, Quincy Point Power Plant of the 
Old Colony Street Railway Company. 



designers. It became evident at once that the old types that were 
capable of creating a 26" or 27" vacuum would require considerable 
modification to maintain a vacuum of 28" or 29". The principal 
improvement adopted by practically all manufacturers has been to 
apply a separate dry vacuum pump for the removal of air and non- 
condensable vapors. 

Surface Condensers. — Fig. 227 shows the arrangement advocated 
by the H. R. Worthington Company. The equipment comprises a 
surface condenser, a steam-driven centrifugal pump for circulating 
the cooling water, a steam-driven rotative dry-air pump, and a motor- 



442 



STEAM POWER PLANT ENGINEERING 




CONDENSERS 



443 



driven centrifugal hot-well pump. The surface condenser is piped 
direct to the turbine exhaust, only a corrugated copper expansion joint 
and a tee intervening. A tubular water vapor cooler, which is in reality 
a small surface condenser, is inserted in the circulating water line 
between the pump suction and condenser, and serves to arrest all the 
condensable vapor and thus reduces the volume to be handled by the 
air pump. All condensation, including that from the air cooler, collects 
in the hot well, from which it is pumped by a motor-driven circulating 
pump direct to heater or boiler. Cooling water is handled by a cen- 
trifugal pump having both suction and delivery pipes water-sealed, so 
that the work done by the pump is virtually that of overcoming the 



Exhaust Steam Inlet 



Water Outlet 




"Fig. 228. High- Vacuum System, C. H. Wheeler Co. 



fluid friction in the condenser and piping. All valves and stuffing 
boxes are water-sealed to prevent any possible leakage of air, and the 
condenser pump cylinder is especially designed to avoid vapor binding. 
This makes it possible to maintain a vacuum of one-half pound absolute 
with cooling water at 60 degrees F. In the high-vacuum condenser 
installation of the Chicago Edison Company the hot-well pump, dry-air 
pump, and the circulating pump are direct connected to a single- 
cylinder Corliss engine. 

Fig. 228 shows the general arrangement of the C. H. Wheeler Com- 
pany's high-vacuum condensing outfit. The condensing chamber is 
shown in section in Fig. 214 and is described in paragraph 219. The 
wet-air pump is illustrated in Fig. 287 and is described in paragraph 288. 
No dry-air pump is needed, and the makers, guarantee a vacuum within 



4U 



STEAM POWER PLANT ENGINEERING 



two inches of absolute under full-load conditions of operating steam 
turbines. 

Fig. 229 shows a section through a Parsons " vacuum augmenter v 
for increasing the vacuum in a surface condenser. A pipe is led from 
the bottom of the main condenser to an auxiliary or augmenter having 
about one-twentieth of the cooling surface of the main condenser. At 
the point indicated a small steam jet is provided which acts as an ejector 
and draws out the air and vapor from the condenser and delivers it to 
the air pump. The water seal prevents the air and vapor from returning 
to the condenser. With this arrangement, according to tests conducted 
by Mr. Parsons, if there is a vacuum of 21\ or 28 inches in the condenser, 
there may be only 26 at the air pump, which, therefore, may be of small 




Fig. 229. Parsons Vacuum Augmenter. 



size, the jet compressing the air and vapor from the condenser to about 
one-half of its original volume. The steam jet uses about one and one- 
half per cent of the steam used by the turbine at full load. 

Jet Condensers. — Fig. 229a gives the general details of a Westing- 
house-Leblanc multi-jet condenser which, under commercial conditions, 
has realized vacua within 99 per cent of the ideal. The most striking 
feature of this system lies in its compactness and simplicity, a 1500- 
kw equipment being less than 9 feet in height. Referring to Fig. 
229a, exhaust steam enters the condenser chamber at the upper left- 
hand opening and meets the cooling water as it is forced through spray 
nozzle C. The condensed steam and injection water fall to the bottom 
of the condenser and are removed by centrifugal pump M. The non- 
condensable vapors are withdrawn by valveless rotary air pumps P, 
through suction opening 0. Referring to section N-N through the 



CONDENSERS 



445 



air pump it will be seen that this pump consists primarily of a reverse 
Pelton turbine wheel in conjunction with an ejector. Sealing water is 
introduced through the branch indicated by dotted outline, into the 
central chamber G, from which it passes through port H. It is then 
caught up by the blades P of the Pelton wheel, which is rotated at a 
suitable speed, and ejected into the discharge cone in the form of thin 




SECTION M.-M, 
THROUGH WATER PUMP. 



Fig. 229a. Westinghouse-Leblanc Multi-jet High- vacuum Condenser System. 



sheets having a high velocity. These sheets of water meet the sides 
of the discharge cone and thus form a series of water pistons, each of 
which entraps a small pocket of air and forces it out against the atmos- 
pheric pressure. In passing through the air pump the sealing water 
receives practically no increase in temperature, hence the same water 
may be used over and over again. The air pump rotor and main pump 
runner are enclosed in a common casing mounted on the same shaft. 



446 



STEAM POWER PLANT ENGINEERING 



This arrangement makes the plant very compact and requires the use 
of only one motor to drive both pumps. There is a clear passage throurh 
the condenser and pump, so that should the pump stop for any reason 



Exhaust from Turbine 




Fig. 229b. Tomlinson Type C High- vacuum Jet Condenser. 

air rushes into the condenser through the air pump and immediately 
breaks the vacuum. In starting up the condenser, steam is turned into 
auxiliary nozzle L, section N-N , for a few moments, thus creating suf- 
ficient vacuum to start the regular flow of water through the air pump. 




Fig. 229c. Section through Condensing Chamber of Kbrting Multi-jet Condenser. 
Chamber Capable of Maintaining a Vacuum of 95 per Cent of the Ideal without 
the Use of Air Pumps. 



The pumps require from H to 3 per cent of the power generated by 
the main engines. Fig. 223 shows an application of a Westinghouse- 
Leblanc condenser to a Curtis turbine. 



CONDENSERS 447 

229. Power Consumption of Condenser Auxiliaries. — In estimating 
the cost of producing vacua with the different types of auxiliaries, 
steam driven, electrically driven, or belted, the power consumption is 
most conveniently expressed in terms of the equivalent heat consumption 
of the auxiliary in question and not the indicated or developed power. 
For example, suppose a power plant has a number of 1200-LH.P. 
engines direct connected to 800 -kilowatt generators and that the engines 
use 20 pounds of steam per LH.P. hour at rated load; furthermore 
suppose the engine driving the air pump (jet condenser) to indicate 
24 horse power. Now, it is manifestly incorrect to say that the power 

24 

consumption of the air pump is equivalent to = 2 per cent of the 

main engine power unless the engine driving the air pump uses 20 
pounds of steam per I.H.P. As a matter of fact the small engine proba- 
bly uses 30 to 40 pounds or more of steam per I.H.P. hour, and the true 
power consumption is 

24 X 30 



1200 X 20 



= 3 per cent, or more. 



If the exhaust steam is piped to the condenser, then all of this 3 per 
cent or more should be charged against the condenser; if the steam is 
piped to a heater, then only the difference between the heat enter- 
ing the small engine and that given up to the feed water should be 
charged against it. For example, suppose the engine in the preceding 
examples uses 30 pounds of steam per I.H.P. hour when running 
condensing and 40 pounds when operating non-condensing. Let 
the initial steam pressure be 150 pounds and feed-water temperature 
120 degrees F. when the air pump is running condensing. If the boiler 
feed is not taken from the hot well, the heat in the exhaust steam is 
lost so far as the economy of the plant is concerned, and the heat con- 
sumption per I.H.P. hour is 30(1193.6 - (120 - 32)} = 33,168 B.T.U. 
This represents the cost, in heat units, of producing the vacuum, and 
is equivalent to 3 per cent of the main engine output. 

If the air pump runs non-condensing and the exhaust steam is piped 
to the heater, each pound of exhaust steam gives up approximately 
950 B.T.U. per hour to the feed water and the temperature of the 
latter is raised from 120 to 180 degrees F. The heat entering the air 
pump is 40(1193.6 -(120 -32)} = 44,224 B.T.U. per I.H.P. hour. 
But 40 X 950 = 38,000 B.T.U. are returned to the feed water. Hence 
44,424 — 38,000 = 6224 is the net heat consumption of the air pumps 
per I.H.P. hour. This corresponds to approximately 0.55 per cent of 
the main engine output. 



448 



STEAM POWER PLANT ENGINEERING 



In the preceding example suppose the air pump to be motor driven 
and that it requires 20 electrical horse power per hour. This will be 

20 

the equivalent of — = 26.2 I.H.P. of the main engine on the 

0.85 X 0.90 & 

assumption that the efficiency of the small motor is 85 per cent and that of 

the engine and generator combined 90 per cent. The power required by 

the air pump will be 26.2 4- 1200 = 2.2 per cent of the total output. 



8 






















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RELATION Or POWER CONSUMPTION 

OF AUXILIARIES TO STATION 

OUTPUT 

Citizens Light, Heat and Power Co. 

Johnsto-wn, Pa. 






7 






\ 








































6 








\ 






















\ 
















5 










\ 












Auxiliary Input % of Turbine Output I.H.P. 
" » » Generator " E.H.P. 
















\ 








_- 






1 


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Full 


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100 



300 400 500 
Load E.H.P. 

Fig. 230. 



700 



In practice the auxiliaries use the equivalent of from 1 to 15 per cent 
of the main engine or turbine steam, depending upon the size of the plant, 
character and number of auxiliaries, and the conditions of operation. 



11 


























1 1 ! 1 1 1 1 1 1 1 1 1 1 1 


































POWER CONSUMPTION OF AUXILIARIES 








12 


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2000 K. W. Curtis Turbine 
La. Purchase Exposition 


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Fig. 231. 



2500 



Table 55 gives the power consumption of the condenser auxiliaries 
in a number of installations. Fig. 230 shows the relation between the 
power consumption of the auxiliaries and the total output of the 



CONDENSERS 



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450 STEAM POWER PLANT ENGINEERING 

station at different loads for a Parsons steam turbine installation, and 
Fig. 231 shows a similar relation for a 2000-kilowatt Curtis turbine. 
(J. R. Bibbins, Power, January, 1905.) 

Steam vs. Electric Auxiliaries : Engr. U.S., 1902, p. 113; Power, Feb., 1905, p. 90, 
Sept., 1906, p. 502; St. Ry. Review, March, 1898, p. 184, July, 1899, p. 458. 
Centralized Control of Auxiliaries : Engr. U.S., Nov. 1, 1906, p. 782. 

230. Cost of Condensers. — The following figures give an idea of 
the relative costs of the different types of condensers and auxiliaries 
for a 1000-I.H.P. plant using 20 pounds of steam per I.H.P. hour at 
rated load, or a total of 20,000 pounds per hour. Vacuum to be 
maintained, 26 inches, unless otherwise stated; temperature of cooling 
water, 70 degrees F.; hot-well temperature, 105 to 120 degrees F.; dis- 
tance between engine exhaust opening and mean level of intake well, 
10 feet. 

Siphon Condensers. 

1 16" siphon condenser with 6" centrifugal pump driven by 6" by 6" 

vertical engine $800 

Jet Condensers. 

1 14" by 22" by 24" jet condenser with single horizontal direct-acting 

pump 1335 

1 16" by 24" by 18" jet condenser with single vertical direct-acting 

pump 1620 

1 14" by 24" by 18" jet condenser with single vertical fly wheel vacuum 

pump 1770 

1 12" by 17" by 22" by 25" jet condenser, single horizontal direct- 
acting compound pump 2200 

Barometric Condensers. 

1 barometric condenser; 10" by 16" by 12" horizontal single-cylinder 
rotative dry-air pump; 8" horizontal volute centrifugal pump direct 
connected to 23-horse-power high-speed engine 2500 

1 barometric condenser; 16" by 16" dry-air pump direct connected to 
9" by 16" steam engine; positive rotary pump, for circulating 

cooling water, belted to above engine 4300 

Surface Condensers. 

1 surface condenser, 1025 square feet cooling surface, mounted over 

1\" by 14" by 14" by 12" combined air and circulating pump 2100 

1 surface condenser, 1025 square feet cooling surface, with 1\" by 12" 
by 12" horizontal air pump, direct acting, and 6" centrifugal pump 
driven by 5" by 5" engine 2300 

1 surface condenser, 1025 square feet cooling surface; 5" by 12" by 
10" Edwards single-cylinder air pump and 6" centrifugal pump 
driven by a 5" by 5" engine ; maximum 28", referred to 30" barometer . . 2850 

1 surface condenser, 1025 square feet cooling surface; 6" by 8" rotative 
dry-air pump; 6" by 6" Edwards wet-air pump and 6" centrifugal 
pump driven by 5" by 5" engine; maximum vacuum 29", referred to 
30" barometer (temp, cooling water 50 deg. F.) 3500 



CONDENSERS 



451 



In general the cost of complete condensing equipments installed and 
ready for operation will approximate as follows: 

Cost per Kilowatt of Main 
Generating Unit. 

Siphon condensers without air pump $2.00 to $ 3.00 

Jet condensers 3.00 to 4.50 

Barometric condensers with dry-air pump 4.00 to 6.00 

Surface condensers for 26-inch vacuum 3.50 to 5.00 

High- vacuum surface condensers 3.50 to 10.00 















1 1 II 1 1 1 1 1 II 1 1 1 1 
























RELATIVE COST OP HIGH VACUUM 
Condens'ng Apparatus 


























































| 










































































































> 






































- 












28 




















































be 

.a 

-3 
a 

o- 

-"3 










































































y 
























































Comprising 
^Surface Condeasers>, 
Dry Air Pumps, 
Circulating Pumps, 
Hot Well Pumps. 




27 






















































































55 






























26 








/ 
























Piping, Valves, etc. 








/ 




















































/ 




















































/ 


















































OS, 


/ 








r 


iceut of( 


'08 


t C 


f Appa 


ra 


US f( 


'' 


6 [Vacuum 









100 



120 140 160 

Fig. 232. 



200 



The curve in Fig. 232 shows the relative costs of complete surface 
condensing plants for steam turbines to maintain the vacua indicated. 

It will be noted how much more 
expensive a high-vacuum plant is 
than one designed for moderate 
vacua. Thus a 27- inch plant 
costs 25 per cent more than a 
26 -inch plant, and a 28.5 -inch 
plant costs twice as much. (J. R. 
Bibbins, Power, January, 1905.) 
The real cost of a condensing 
plant, however, is not limited to 
the cost of condensing auxiliaries 
and piping, but should include 
all other costs necessitated by 
the use of the condensing plant, including cost of extra building space, 
foundations and the like, and the attending fixed charges. 

231. Most Economical Vacuum.* — The load factor, or the ratio 
of the actual yearly load to the rated yearly capacity, has a marked 
influence on the degree of vacuum best suited for a given installation, 
since the fixed charges go on whether the plant is running or not, while 
the gain due to the higher vacuum is realized only when the engines 
are operating. The higher the load factor the greater is the amount of 
power produced, the longer does the apparatus operate at best effi- 
ciency, the lower the ratio of fixed charges to total operating expenses, 
and consequently the lower the cost of power per unit. 

The load factor for electric-lighting stations is invariably low and 
seldom exceeds 25 per cent, with an average not far from 18 per cent. 
In street-railway work it is higher and averages about 30 per cent. In 
manufacturing plants the load factor varies considerably, but as a rule 
is somewhat higher than in either of the above cases. Tables 56 and 
57 {Power, December, 1906, p. 769) show the most economical vacua 
for different load factors for plants of 1000 kilowatts capacity with 
* See also, Elec'n, Lond , Jan. 14, 1910. 



452 



STEAM POWER PLANT ENGINEERING 



conditions as stated. From the tables it would seem at first glance 
that, except where coal is expensive, all the plants with low factors, 
10 per cent and under, ought to be run non-condensing. This is true 
for " one-engine " installations, but not necessarily so where there 
are a number of engines or turbines. In the latter case higher 
economy may be effected by providing only a portion of the engines 
with condensing equipment. The engine carrying the continuous or 
day load should operate condensing, and the non-condensing engine 
should carry the peak load. In order that any of the units may be 
used for the day work, all engines could be connected to the condenser, 
but only those carrying the day load should be operated condensing. 
Each installation, of course, must be considered separately and due 
weight given to the various factors entering into the problem. For an 
excellent article on the subject see " Condensers for Steam Engines and 
Turbines," Power, December, 1906, p. 769, and the Engineer, London, 
April 13,' 1906, p. 381. 

232. Choice of Condensers. — The proper selection of a condenser 
for a proposed installation depends upon the conditions under which 
the plant is to be operated. When there is a plentiful and cheap sup- 
ply of good condensing water suitable for boiler feed, and extremely high 
vacua are not essential, some type of jet condenser will generally be 
found most desirable. If overhead room permits, a siphon or baro- 
metric condenser will probably be most suitable and least expensive. 



TABLE 56. 
MOST ECONOMICAL VACUUM FOR STEAM TURBINES. 

Vacuum referred to 30-Inch Barometer. 





Cost of Coal, Dollars per Ton. 


Load Factor, 
per Cent. 


$1.50 


$2.00 


$2.50 


$3.00 


$3.50 




A 


B 


A 


B 


A 


B 


A 


B 


A 


B 


5 

10 


N.C. 

20 

24 

26.5 

27.5 

28 


N.C. 

N.C. 

17 

20 

24 

27.6 


N.C. 
23 

26.5 
27.3 

27.8 
28.2 


N.C. 

N.C. 

20 

23 

27 

27.9 


18 

25 

27 

27.6 

28 

28.3 


N.C. 

N.C. 

22 

25.5 

27.6 

28 


20 

26.5 

27.5 

27.8 

28.1 

28.4 


N.C. 

20 

24 

27 

27.8 

28 


22 

27 

27.7 

27.9 

28.2 

28.5 


N.C. 
22 


15 


25.8 


20 


27.5 


30 


28 


50 


28 















A. Surface-condensing plant; cost $6 per kilowatt of main generator. Fixed charges 12 per cent. 
Cost of water not included. Rated capacity of generator, 1000 kilowatts. 

B. Surface-condensing plant, including cooling towers and extra cost of land, etc.; cost $10 
per kilowatt for 26-inch plant, increasing to $14 per kilowatt for 28.5-inch plant. Fixed charges 
12 per cent. No charge for water. Rated capacity of generator, 1000 kilowatts. 



CONDENSERS 



453 



TABLE 57. 

MOST ECONOMICAL VACUUM FOR RECIPROCATING ENGINES. 

Vacuum referred to 30-Inch Barometer. 





Cost of Coal, Dollars per Ton. 


Load Factor, 
per Cent. 


$1.50 


$2.00 


$2.50 


$3.00 


$3.50 




A 


B 


A 


B 


A 


B 


A 


B 


A 


B 


10 

15 


N.C. 

16 

22.5 
24 
25.5 


N.C. 

N.C. 
N.C. 

16 

22 


15 

20 

23 

24.5 

26.7 


N.C. 

N.C. 

N.C. 

21 

23.5 


18 
22 

23.5 
25.5 
27.2 


N.C. 

N.C. 

20 

22 

23.5 


20 

22.5 

24.5 

26.4 

27.5 


N.C. 

16 
21 
23 
26.3 


22 
24 
25 

26.8 

27.7 


N.C. 
20 


20 


22 


30 


24 


50 


27 



A. Surface-condensing plant; cost $7 per kilowatt of main generator. Fixed charge 12 per cent. 
Cost of water not included. Rated capacity of generator, 1000 kilowatts. 

B. Surface-condensing plant, including cooling towers and extra cost of land, etc.; cost $11 per 
kilowatt for 26-inch plant, increasing to $13 per kilowatt for 27.5-inch plant. Other conditions 
as in A. 

Where there is a plentiful supply of good water for boiler feed but 
the water which must be used for cooling purposes is very dirty the 
siphon condenser is preferable to the barometric form. A surface con- 
denser may be used in the latter case if the condensing water is not so 
dirty as to seriously impair the efficiency by coating the tubes with 
sediment, and boiler feed water is scarce. 

The air-cooled surface condenser is employed only where water of any 
kind is scarce. 

For very high vacua in connection with steam turbine work the sur- 
face condenser is almost universally adopted, although the barometric 
condenser in connection with dry-air pumps is finding favor with many 
engineers. 

In selecting the type of condenser and auxiliaries due weight 
must be given to the load factor, cost of coal, water, land, building, 
interest, depreciation and the like, as outlined in the preceding 
paragraph. 

233. Water-Cooling Systems. — When an ample supply of cooling 
water is unobtainable for natural or economic reasons, the circulating 
water may be used over and over again by employing suitable cooling 
devices. The three most common in practice are 

1. The simple cooling pond or tank. 

2. The spray fountain. 

3. The cooling tower. 



454 STEAM POWER PLANT ENGINEERING 

233a. Cooling Pond. — The water is cooled partly by radiation and 
conduction but principally by evaporation. The air is seldom saturated 
normally, and its capacity for absorbing moisture is increased on account 
of its temperature being raised by contact with the warm water and 
by radiation. The cooling action is independent of the depth of water 
and varies directly as the surface, the amount of heat dissipated for 
each square foot depending upon the temperature of the water, the rela- 
tive humidity, and the velocity of the air currents. Results of tests are 
very discordant. 

Box in his treatise on Heat states that the pond surface should 
approximate 210 square feet per nominal horse power for an engine 
working twenty-four hours a day. (Treatise on Heat, Box, p. 152.) 

If the engine works only twelve hours per day, the area may be reduced 
to 105 square feet per horse power, because the water will cool during 
the night, but in that case the depth should be such as to give a capacity 
of 300 cubic feet per horse power. These figures are based on a reduc- 
tion in temperature of 122 to 82 degrees F., with air at 52 degrees F. 
and humidity 85 per cent, the steam consumption per nominal horse 
power being taken at 62.5 pounds. 

Box gives the following formula for the rate of evaporation in per- 
fectly calm air: 

E = (243 + 3.7 (V-v), (95) 

in which 

E = evaporation in grains per square foot per hour. 
t = temperature of the water, degrees F. 

V = maximum vapor tension in inches of mercury at temperature t. 
v = actual vapor tension. 

Evaporation is greatly affected by the force of the wind and varies 
from 2 to 12 times the amount determined from equation (95). 

Example: How many pounds of water will be evaporated per square 
foot per hour from a pond with the temperature of the water and air 
80 degrees F.; air perfectly calm; barometric pressure 29.5 inches and 
relative humidity 70 per cent? 

The maximum vapor tension at temperature of 80 degrees is 1.02 
inches of mercury. The actual vapor tension will be 

1.02 X .70 ( = relative humidity) = .714. 
Substitute these values in (95). 

E = (243 + 3.7 X 80) (1.02-0.714) 
= 165 grains per square foot per hour. 
= .023 pound per square foot per hour. 



CONDENSERS 



455 



If the temperature of the water were 130 degrees F. and that of the 
surrounding air 80 degrees F., humidity 70 per cent, the evaporation 
would be 

E = (243 + 3.7 X 80) (4.5-0.714) 
= 2040 grains per square foot per hour 
= 0.291 pound per square foot per hour. 

Here 4.5 = maximum vapor tension, corresponding to a temperature 
of 130 degrees. 

233b. Spray Fountain. — From equation (95) we see that even 
under the most favorable circumstances an enormous pond surface is 
necessary. To facilitate evaporation with a view toward reducing the 
size of the pond, the hot circulating water is sometimes distributed 
through pipes and discharged through nozzles, falling to the surface of 
the pond in a spray. The following gives some interesting data con- 




100 



9 GO 



2.40 









/ 


\ 


1 | 1 1 

Water at Nozzles 


"\ 






J 


I 


[— 


^y 








\-T 


^J 


/ 


iN 


f 


Water in Pond 


A 


r 


\ , 


/ 


\ 


r 


1 


\> 






U' 


w 


/ 


\ 


7 










iL 




A 




/ — 


J.' 


ir 




1/ 


\, 


j 


\ 


/ 


/ 








y 




\ 


z_ 













L 


t 


10 13 1G 1 


9 22 25 28 




L i 


t 


r i 


13 1G 19 22 25 28 


§ 90 

^ 80 


A 






A 


fs 




|100 
*80 

J» 

W 60 

CD 
I 50 

« 40 








$T 




i 








h 8U 
*70 


Jh 


r\ 


A 




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A 




I 


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v. 


y ' 


/ 


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/ 


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i 




a 00 

3 
















N 






\ 


f 


v 




\ 






s 


V 




> 








v 














J 


V 




V> 










3 i0 

(2 30 


i 






V 














V 

















1 4 7 10 13 16 19 22 25 28 
Day o£ the Month September, 1901 



4 7 10 13 16 19 22 25 
Day -of the Month January, 1905 



Fig. 233. Curves Showing Performance of Spray Fountain; Chattanooga Electric 
Company's Power Plant. 



cerning the spray fountain installation at the power plant of the Chat- 
tanooga Electric Company, Chattanooga, Tenn. {Street Railway Review, 
March 15, 1905.) 

Adjoining the power house a pond 150 x 300 feet was excavated to 
a depth of 4 feet, the level of the water being 8 feet below the condensers. 



456 STEAM POWER PLANT ENGINEERING 

Circulating water returned from the condensers is distributed through 
a set of pipes provided with 42 nozzles through which the water is dis- 
charged upwards. The rectangle denned by the center lines of the 
outermost pipes is 98 feet by 125 feet. The pipes are supported on 
brick piers spaced at intervals of about 20 feet in each direction. The 
discharge opening of the nozzles is 1J inches in diameter, and the interior 
is provided with a spiral core so that in its passage the water is given 
a rotary motion, the effect of which is to greatly increase the spraying 
action. The nozzles, except on the extreme outer lines of piping, are 
placed in pairs with the axes in a vertical plane at right angles to the 
center line of the supply pipe, the axis of each nozzle making an angle 
of 30 degrees with a vertical plane through the center of the supply 
pipe. The effect of each pair of nozzles is to throw a mass of spray to 
the height of about 15 feet, which in falling covers an area of 15 x 30 
feet. 

A dike extending nearly across the pond near one end provides a 
canal through which the water is conducted to the suction chamber, the 
object being to draw the supply from distant parts of the pond to give 
greater time for cooling. The " make-up " water is supplied by wells. 
The operation of the cooling pond for a warm month and for a cold 
month is shown in Fig. 233. Readings were taken at three-hour inter- 
vals. The pond supplies the circulating water for three 2000-square- 
feet Worthington surface condensers. 

234. Cooling Towers. — A cooling tower consists of a wooden or 
sheet-iron housing open at the top and bottom and so arranged 
that the heated cooling water may be elevated to the top and dis- 
tributed in such a manner that it falls in thin sheets or sprays into 
a reservoir at the bottom, air at the same time being drawn in at the 
bottom by natural draft or forced in by a fan. The water gives up its 
heat to the ascending current of air by evaporation and conduction, the 
latter, however, being a relatively small factor. If the air supply is 
dependent entirely upon convection, the system is known as the natural- 
draft or flue cooling tower; if the air is forced into the tower by fans, it 
is called a fan cooling tower. The different types vary principally in 
the method of water distribution. Fig. 234 illustrates the Barnard 
cooling tower, in which the falling water is broken up by vertically 
suspended galvanized iron wire cloth mats, causing it to trickle in thin 
sheets to the bottom. A similar result is brought about in the Worth- 
ington tower, Fig. 235, by pieces of terra-cotta pipe 6 inches in diame- 
ter and two feet long placed on ends in rows. In the Alberger cooling 
tower the water trickles down the sides of swamp-cypress boards 
arranged in honeycomb fashion. In the Jennison cooling tower the 



CONDENSERS 



457 



water is divided into a rain of drops, constantly retarded in their fall by 
a series of perforated 4x4 inch galvanized iron trays arranged in 
horizontal rows and staggered vertically. 

With the best forms of cooling towers, under average conditions, the 



DISTRIBUTING 
TROUGH 




DISCHARGE 
FROM 
TOWER 



Fig. 234. Barnard- Wheeler Cooling Tower. 



temperature of the circulating water may readily be reduced from 40 
to 50 degrees with a loss not exceeding 3 or 4 per cent of the total 
quantity of water passing through the tower. The power consumed 
by the fan in a forced-draft apparatus averages 2 per cent of that 



458 



STEAM POWER PLANT ENGINEERING 



developed by the main engines, for the maximum requirements during 
summer months, and 1 J per cent during the winter. 

The location of the tower may be on the engine-room floor, on top of 



TOWD\ 




< ■ CC « HOT WATER. 



COLD WATER,. 



Suction tank 

Fig. 235. Worthington Cooling Tower. 



the building, or in the yard, the latter being the most adaptable. It 
may be any reasonable distance from the engine and condenser. 
Fig. 236 shows a typical installation of Alberger condenser and cooling 
towers. 



COiNDENSERS 



459 




460 STEAM POWER PLANT ENGINEERING 



235. Parallel Comparison of Fan and Natural-Draft Cooling Towers. 

Fan. Natural Draft. 

Size. 

Small, the forced draft providing Large, draft being necessarily small, 
sufficient air velocity to effect evapo- a larger area must be provided to 
ration. perform same work. 

Height limited, because loss from back Height is an advantage because the 
pressure increases with the height. tower operates on the principles of a 

Tower usually short and of large area. chimney. 

Power Consumption. 

One per cent of station output and None. 
upwards, depending upon the type 
of auxiliaries and the conditions of 
operation. 

Location. 

Inside or outside. Can operate in any Outside only, unless exceptionally good 

location where sufficient head room draft is obtainable. 

and air supply are available. Preferably in the open where advan- 

Especially adapted to inconvenient lo- tage may be taken of prevailing 

cations, as roofs, upper decks, boiler winds. 

floors, etc. 

Conditions of Atmosphere. 

Comparatively little affected by tern- Largely affected by temperature and 
perature, considerably by humidity, humidity and wind. Draft increased 
and none by winds. by steady winds. 

Conditions of Operation. 

More especially adapted for heavy con- Especially adapted for light summer 
tinuous duty the year round, as in and heavy winter duty, as in electric- 
rail-plants or mills. lighting plants. 

First Cost and Cost of Operation. 

First cost greater on account of First cost small by reason of simplicity 

mechanical construction and neces- and construction. 

sary auxiliaries. First cost largely dependent upon ma- 
Cost of operation dependent upon type terials used in interior construction. 

of auxiliary and conditions of oper- Cost of operation limited to fixed 

ation. charges. 

236. Water-cooling Calculations. — Air is said to be completely 
saturated when it contains all the water vapor it can hold without 






CONDENSERS 461 

causing precipitation. If the vapor content is less than that corre- 
sponding to complete saturation the air will tend to become saturated 
by absorbing moisture from surrounding objects. The drier the air 
the greater will be its affinity for moisture. The necessary latent heat 
for vaporization is supplied directly by the water producing the vapor 
or by the surrounding objects in contact with the water. Thus, in the 
open cooling-tower the water vapor is absorbed from the circulating 
water, and the heat necessary to effect this vaporization is given up by 
the water, with a resultant reduction in temperature of the water itself; 
and in the evaporative surface condenser the vapor is absorbed from 
the water spray in contact with the tubes, the heat required to effect 
this vaporization being given up by the steam within the condenser 
chambers, resulting in condensation of the steam. If the air coming 
in contact with the water is very dry and at a high temperature the 
vaporization of the water may be rapid enough to cool the remaining 
water to a temperature much lower than that of the air. In this case 
practically all of the cooling is effected by evaporation. But when the 
air is at a low temperature and high relative humidity a considerable 
amount of heat may be carried away by the air by conduction. The 
quantity of air and water necessary to produce a given cooling effect 
may be determined as follows: 

Let H = total amount of heat to be abstracted, B.T.U. per hour. 

W = weight of water to be cooled, lbs. per hour. 

t e = temperature of water entering cooling device. 

ti = temperature of water leaving cooling device. 

t = temperature of air entering cooling device, °F.;T = £ + 460. 

U = temperature of- air leaving cooling device, ° F. ; T 2 = t 2 + 460. 

p = ordinary atmospheric pressure = 29.92 in. of mercury. 

p a = observed atmospheric pressure, in. of mercury. 

p = elastic force of vapor at temperature t , in. of mercury. 

p 2 = elastic force of vapor at temperature t 2 , in. of mercury. 

V = volume of air entering the cooling device, cu. ft. per hour, 
atmospheric conditions. 

V 2 = volume of air discharged from the cooling device at tem- 
perature t. 
d = density of dry air, at pressure p and temperature t . 

h = weight of moisture in 1 cu. ft. of saturated air at tempera- 
ture t , pounds. 

h 2 = weight of moisture in 1 cu. ft. of saturated air at tempera- 
ture t 2 , pounds. 

z = relative humidity of the air entering the cooling device. 

z 2 ==Telative humidity of the air leaving the cooling device. 



462 



STEAM POWER PLANT ENGINEERING 



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CONDENSERS 463 

C = specific heat of dry air at constant pressure, = 0.2375. 
S = mean specific heat of water vapor at temperature U. 
r 2 = heat of vaporization at temperature U. 

The pressure p 3 of the dry air in atmospheric air entering the cooling 
device is Pi = Va _ ^ (96) 

The pressure p 3 of the dry air leaving the cooling device is 

Vs = Va- Vi z 2- (96a) 

The weight w of dry air entering the cooling device under atmospheric 
temperature and pressure, lbs. per hour, is 

w = ^dV . (97) 

J05- 

The weight w of moisture carried into the cooling device by the air, 
lbs. per hour, is Wq = h z V . (97a) 

The volume of air leaving the cooling device is 

V,-V,2i.%?- (98) 

Ih J 

The weight w 2 of moisture carried away by the air discharged from the 
cooling device, lbs. per hour, is 

w 2 = h 2 z 2 V . (98a) 

The weight of circulating water w 3 absorbed by the air in passing through 
the cooling device, lbs. per hour, is 

w 3 = w 2 — w . (98b) 

The ftea£ // to be abstracted from the circulating water, B.T.U. per 
hour, is H = W(t l -t e ). (99) 

The heat is dissipated, by the cooling process, in raising the tempera- 
ture of the air and vapor entering the cooling device from t to t 2 (by- 
conduction) and in evaporating the moisture absorbed by the air in pass- 
ing through the apparatus (by evaporation). 

The heat H a required to raise the temperature of dry air from t to t 2 , 
B.T.U. per hour, is H a = Cw (t 2 - t ). (100) 

The heat H s required to superheat the water vapor entering the cool- 
ing device from temperature t to t 2 , B.T.U. per hour, is 

H s = w S(t 2 -t ). (100a) 

The heat H c abstracted by conduction from the circulating water, 

B.T.U. per hour, is H c = H a + H s . (100b) 

The heat H e abstracted from the circulating water by evaporation, 

B.T.U. per hour, is H = wr (101) 



464 STEAM POWER PLANT ENGINEERING 

Though the process of evaporation is practically continued through 
the whole range in the cooling device, we are justified in using the heat 
of vaporization at the highest temperature, because the liquid was at 
this temperature entering the cooling device and the vapor is brought 
back to the temperature when leaving it. 

The total heat H c absorbed by the air in passing through the cooling 
device, B.T.U. per hour, is H t = H c + H e . (102) 

Neglecting radiation and other minor losses, the heat H t absorbed 
by the air must be equal to the heat given up by the circulating water, or 

H t = H. (103) 

Example: Determine the quantity of air passing through the cool- 
ing tower per hour and the circulating water lost by evaporation in a 
power plant operating under the following conditions: Engines indicate 
500 H.P. and consume 20 lbs. steam per I.H.P. hour; temperature of 
the injection water, discharge water and outside air, 90, 122 and 72° F., 
respectively; barometer 29.5; relative humidity of air entering and 
leaving tower 70 and 90 per cent respectively; vacuum at condenser 
25 inches. Determine also the weight of water evaporated in per cent 
of that circulated and of the condensed steam. 

In the problem, 



fti = 29.5, 


h = 72, 


t K = 


0.001224, 


t Vo = 0.79, 


* t 2 = 112, 


t K = 


0.003978, 


t P 2 = 2.79, 


t e = 122, 


v= 


0.70, 


d = 0.0747, 


k= 90, 


*2 = 


0.90, 


C= 0.2375, 


S = 0.45, 


r 2 = 


1028.9. 



These values are obtained from Steam Tables and from Air Tables 
(Table 58). 
Substitute these values in equations (96) to (103) thus: 

(96), Vi = 29.5 - 0.79 X 0.7 

= 28.95. 
(96a), p 3 = 29.5 - 2.74 X 0.9 

= 27.03. 

(97), w = gg X 0.0747 V . 

= 0.0722 V . 
(97a), w = 0.001224 X 0.7 V 
= 0.000857 V . 

* By assumption, t 2 being 10 to 20 degrees lower than t e in average practice when 
the range t e — t is greater than 30 degrees. 

t Marks and Davis: the values in Table 58 are Regnault's. 



CONDENSERS 465 

(98) v = ggj5 t 460 + 112 
W), V2 2703 460+ ^^o 

= 1.152; 

that is, each cu. ft. of dry air entering the cooling-tower is increased in 
volume to 1.152 cu. ft. as it leaves. 

(98a), w 2 = 0.003978 X 0.9 X 1.152 V 

= 0.004125 V . 
(98b), w 3 = 0.004125 V - 0.000857 V 
= 0.003268 V . 

The total heat to be abstracted from the steam (see equation (84), 
page 347) is 

H = 500 X 20 (1120.1 - 122 + 32) 
= 10,300,000 B.T.U. per hour. 
(99), But W (122 - 90) = 10,300,000, 
from which W = 322,000 pounds per hour. 

(100), H a = 0.2375 X 0.0722 V (112 - 72) 

= 0.6865 V . 
(100a), H s = 0.000857 V X 0.45 (112 - 72) 

= 0.001543 V 
(100b), H c = 0.6865 V + 0.001543 V 

= 0.688 V . 
(101), H e = 0.003268 V X 1028.9 
■= 3.365 7 . 
g e _ 3.365 7 
H c 0.688 V ' 

that is, the air removes 4.89 times more heat by evaporation than by con- 
duction under the given conditions. 

(102), H t = 0.688 V + 3.365 V 
= 4.05a v . 

that is, the heat required to superheat the moisture carried into the 
tower by the air is approximately T $o of 1 per cent of the total; hence 
an error as great as 20% in the mean specific heat of the vapor is 
negligible. 

(103), 4.053 V = 10,300,000, 
from which 

V = 2,543,000 cu. ft. of air per hour 

necessary to effect the required cooling 
= 42,300 cu. ft. per minute. 



466 STEAM POWER PLANT ENGINEERING 

From (98b) 



Substitute 



0.003268 V, 



V = 2,543,000 in above equation. 

w 3 = 0.003268 X 2,543,000 

= 8320 lbs., or the weight of circulating water 
carried away per hour. 

m, 8320 



W 322,000 



= .0258; 



that is, 2.58 per cent of the circulating water is carried away by the air 
in effecting the necessary cooling. 

*> _ 8320 
20 X 500 10,000 ' 

that is, the equivalent of 83.2 per cent of the steam used by the engines 
is evaporated in the cooling tower, or the make-up water is more than 
supplied by the condensed steam. 

Example: Evaporation Surface Condenser. — How many cubic feet 
of air and how many pounds of water spray must be forced through an 
evaporative surface condenser of the fan type in order to condense 
1000 pounds of steam per hour and maintain a vacuum of 25 inches, 
barometer 29? (Atmospheric air 80° F., relative humidity 70%.) 
The air and vapor issue from the discharge pipe under pressure of four 
inches of water, temperature 120° F., relative humidity 98%. 

The absolute pressure in the condenser is 29.0 — 25.0 = 4 inches of 
mercury. 

The total heat to be withdrawn in order to cool and condense 1000 lbs. 
of steam per hour at absolute pressure of 4 inches to 120° F. is 

1000 [1114.8 - (120. + 32)] = 1,026,000 B.T.U. 

Neglecting radiation and leakage losses, this is the heat to be ab- 
stracted per hour by the air and water spray. 

The pressure of the dry air in the mixture entering the condenser is, 
equation (96), 

p 1 = 29.0 - 0.7 X 1.029. 
= 28.28. 
The pressure of dry air in the mixture leaving the condenser is, 

equation (96a), 

p 3 = (29.0 + 0.294) - 0.98 X 3.438 

= 25.925 

(0.294 is the value in inches of mercury of four inches of water-fan 
pressure). 



1 



CONDENSERS 467 

Let V = volume of atmospheric air entering the condenser. The vol- 
ume leaving the condenser will be, equation (98), 
_ 28.280 460 + 120 _ 
V °~ 25.925' 460+80 " ' V °' 
The weight of vapor in the condenser discharge is, equation (98a), 
w 2 = 1.172 V X 0.004888 X 0.98 
= 0.005615 V lbs. 
The weight of vapor in the mixture entering the condenser is, equa- 
tion (97a), Wq = 0.00157 X 0.7 V 
= 0.001099 V lbs. 
The amount evaporated therefore is 

w 3 = 0.005615 7 - 0.001099 V 
= 0.004516 V lbs. 

The weight of dry air entering the condenser is, equation (97), 
90 90 

W = 29321 007362 ^ 
= 0.06958 y lbs. 
The heat absorbed by the dry air in being heated from 80° to 120° F. 
is, equation (100), 

H = Cw (t 2 - t ) 

= 0.02375 X 0.06958 V (120 - 80) 
= 0.658 V B.T.U. 
Heat required to superheat w lbs. of vapor from 80° to 120° F. is, 
equation (100a), ^ = 0.001099 V X 0.46 (120 - 80) 
= 0.02022 V B.T.U. 
Heat absorbed by the evaporation of w 3 lbs. of water is, equation 
(101), H e = 0.004516 y X 1046.7 

= 4.720 V B.T.U. 
(Here the latent heat is taken at the lower temperature, it being the 
original temperature of the liquid.) 

Total heat absorbed by the entering air and spray is 
H t = 0.658 7 + 4.720 V + 0.020 V 
- 5.398 y . 
But this represents also the heat given up by the steam, or 

5.398 y = 1,026,000. 
From which V = 190,500 cu. ft. of atmospheric air necessary to con- 
dense and cool 1000 pounds of steam under the given conditions. 
The water spray to be injected per hour is 

0.004516 y = 0.004516 X 190,500 = 860 pounds. 



468 STEAM POWER PLANT ENGINEERING 

236a. Hygrometry. — The degree of saturation, or relative humidity, 
is ordinarily determined from the difference in reading of a wet and a dry 
bulb thermometer, thus : If the air is saturated with aqueous vapor no 
evaporation takes place from the wet bulb and the two thermometers 
give identical readings; but if it is unsaturated, evaporation occurs. 
The wet-bulb thermometer is thus cooled and its readings are lower than 
those of the dry bulb. The difference in reading is a function of the 
relative humidity, and the latter may be calculated from the following 
modification of Apjohns' formula: 

If the thermometer reads above 32° F. 



j, / dP \100 

htes { v "~2m) TT (103a) 



If it reads below 32° F. 
h 
in which 

h = relative humidity, per cent. 

d = difference in reading of the wet and dry thermometers, degrees F. 
P = barometric pressure, inches of mercury. 

P w — maximum tension of aqueous vapor corresponding to the 
temperature of the wet thermometer, inches of mercury. 
(This may be taken directly from the Steam Tables.) 
P t = maximum tension of aqueous vapor corresponding to the tem- 
perature of the dry thermometer, inches of mercury. 

Example: Determine the relative humidity when the dry bulb reads 
70° F., wet bulb 60° F., barometer 28.0. 
From the Steam Tables we find 

P w = 0.522; P t = 0.739. 
Whence 

eoo 10 X 28\ 100 
• 522 "-264(r)a739 = 56 - 5percent - 

Tables giving the relative humidity in terms of the temperature 
difference are published in most engineering handbooks and the above 
calculations are unnecessary. These tables, however, are based on a 
fixed barometer pressure, whereas the formula takes the actual pressure 
into consideration. 

237. Test of Cooling Tower (Wheeler Condenser Company), — The 
following gives the results of a test made on the cooling-tower plant of 
the A. F. Brown Company at Elizabethport, N. J. The tower is work- 
ing in connection with a Wheeler surface condenser of 280 square feet 
of cooling surface, mounted over a 10, 12X12 combined air and 
circulating pumn. 



(„. 



! 



CONDENSERS 469 

Observations made on June 24, 1904. 

Temperature of air 81 degrees 

Hygrometer 69 degrees 

Temperature of air at top of lower 89 degrees 

Temperature of water in troughs 105 degrees 

Temperature of water in tank 83 cleg 

Revolutions of fun, 239 r.p.m.. air pressure \ inch water 

Velocity of air out of tower 822 feet per minute 

Gallons of water passing over mats 385 per minute 

Vacuum 20 inches 

Temperature of air-pump discharge 87 degrees 

Observations made June 28, J 904, 9 a.m 

Temperature of air 70 degrees 

Hygrometer 59 degrees 

Temperature of air at top of tower 81 degrees 

Temperature of water in troughs 96 degrees 

Temperature of water in tank 78 degrees 

Revolutions of fan, 232 r.p.m., air pressure jj inch water 

Velocity of air out of tower 080 feet per minute 

Gallons of water passing over mats 400 per minute 

Vacuum 25.5 inches 

Temperature of air-pump discharge 90 degrees 

Observations made June 28, 1904, 3 P.M. 

Temperature of air 74 degrees 

Hygrometer 57 degrees 

Temperature of air at top of tower 83 degrees 

Temperature of water in troughs 99 degrees 

Temperature of water in tank 80 degrees 

Revolutions of fan. 237 r.p.m , air pressure ^ inch water 

Velocity of air out of tower 709 feet per minute 

Gallons of water passing over mats 470 per minute 

Vacuum ... 2.5.5 inches 

Temperature of air-pump discharge 92 degrees 

Observations made June 29, 1904. 

Temperature of air 78 degrees 

Hygrometer 71 degrees 

Temperature of air at top of tower 80 degrees 

Temperature of water in troughs 108 degrees 

Temperature of water in tank 82 degrees 

Revolutions of fan, 211 r.p.m., air pressure § inch 

Velocity of air out of tower 772 feet per minute 

Gallons of water passing over mats 430 per minute 

Vacuum 25.5 inches 

Temperature of air-pump discharge 93 degrees 

Specifications for condensers — See paragraph 414. 

RESULTS OF TEST OF NATURAL-DRAFT TOWER, DETROIT. 
Complete Five-Fifths Surface Installed. 
Proc. A.S.M.E.. Mid-Nov., 1909, p. 1205. 
Engines: Two 400-i.h.p. 300-kw. Macintosh & Seymour tandem-compound 

engines, overhung generators. 
Condensers: Worthington surface (admiralty type) 1000-sq. ft. reciprocating wet- 
air pump and circulating pump 
Tower: Wood-mat construction, 24,500 sq. ft evaporating surface, exclusive 

Of shell 
Test- March 15 to 10, 1901, 4 p.m. to 4 p m , 24 hr. 



470 



STEAM POWER PLANT ENGINEERING 



Weather; 

Load: 
Steam : 



Water: 



Results: 



Cooling: 

Evaporati 
Tower: 



A.M. 

30.22 

18.5 

76 



P.M. AVERAGE. 

30.07; 30.14 30.27 
25; 30 25 

82; 58 72 



Barometer (abs.), min 

Temperature air, deg 

Relative humidity, per cent 

600 kw. max. to 50 kw. min. Average 244.9 kw. 

Engine efficiency = 92.5 = 875 i.h.p. max. Average . .354.8 i.h.p. 

Weight of condensed steam per hr., lb 5910.6 

Temperature exhaust steam, deg. F 134 . 38 

Temperature condensed steam, deg. F 108 . 78 

Weight of steam per hour, max. load, lb 13,500 

Vacuum (abs.) 25 to 19, average about 22 

Vacuum corresponding to temperature exhaust steam. . . 25 

Vacuum possible with good condenser (10 deg. difference) 28 

Circulated per hr., lb 293,536 

Temperature hot well, average, deg. F 87.50 

Temperature cold well, average, deg. F 71.27 

Vaporization loss per hr., lb 5970 

Condenser surface per kw., sq. ft 2.66 

Steam per kw. hr., lb 24 . 3 

Steam per i.h.p. hr., lb 16.66 

Circulating water per lb. of steam, lb 49 . 6 

Steam per sq. ft. condenser surface per hr., lb 3.7 

Circulating water per sq. ft. tower surface, lb 12 

Difference in temperature between exhaust steam and 

discharge, deg. F 47 

Max. 20 deg., min. 3 deg. -5 deg. Average 16.23 

Heat dissipated per hr., B.T.U 4,769,000 

Heat per sq. ft. tower surface, B.T.U 195 

Heat per sq. ft. per 1000 lb. water, B.T.U 0.665 

Circulating water, per cent 2 .03 

Engine steam, per cent 101 

Surface per kw. (average load 245 kw.), sq. ft 100 

Surface per kw. (max. load 600 kw.), sq. ft 40.8 

Surface per 1000 lb. steam max. load, sq. ft 1820 

Surface per 1000 lb. steam average load, sq. ft 4140 

Surface per 1000 lb. circulating water per deg. max. cool- 
ing, sq . ft 4.17 





Temperature, Deg. Fahr. 


Quantities. 


Time. 


Air. 


Hot 
Well.* 


Cold 
Well. 


Water 
Cool- 
ing. 


Total 
Heat 

Head.t 


Tower 

Water, Lb. 

per Hr. 


Heat Dissi- 
pated, B.t.u. 
Lb. per Hr. 


Heat per 
Sq. Ft. Cool- 
ing Surface, 
B.t.ii.perHr. 


Circulating 

Water per 

Sq. Ft., Lb. 

per Hr. 


Load, 
Kw. 


1 


2 


3 


4 


5 


6 


7 


8 


9 


10 


11 


12noon 


34 


102 


89 


13 


68 


375,000 


4,880,000 


332 


25 


270 


1.30 


35 


106.5 


90 


16.5 


71.5 


♦ (375,000 
1370,200 


6,108,000 


415 


24.8 


)315 
1290 


2.30 


35 


106.5 


87.5 


19 


71.5 


375,000 


7,120,000 


484 


25 


315 


3.30 


35 


113 


88.5 


24.5 


78 


375,000 


9,000,000 


613 


25 


350 


4.30 


32.5 


100 


84 


16 


67.5 


399,000 


6,384,000 


434 


26.6 


365 


5.00 


28.5 


103.5 


88 


15.5 


75 


445,500 


6,900,000 


470 


29.7 


485 


6.00 


26 


125 


94 


31 


99 


417,000 


12,930,000 


880 


27.8 


655 


7.00 


24 


121 


94 


27 


97 


427,000 


11,532,000 


785 


27.4 


570 


8.00 


24 


123 


94.5 


28.5 


99 


427,000 


12,174,000 


827 


27.4 


600 



* Assuming a more efficient condenser, say 10 deg. difference, the probable vacuum would be 
26 deg. to 27.5 deg. This condenser actually operated at 40 deg. to 50 deg. difference. 

t Total heat head = air heating + lost head. J Difference due to rapid change in load. 



CHAPTER XII. 

FEED-WATER PURIFIERS AND HEATERS. 

238. General. — All natural waters contain more or less foreign 
matter either in suspension or solution. Waters containing carbonates 
and sulphates of magnesia and lime, soluble salts of silica, iron, and 
alumina, and suspended matter, tend to form scale in the boiler 
and reduce its steam-generating capacity and economy. The loss due 
to this cause is often overestimated but is of secondary importance 
to the danger due to retarded heat transmission which overheats and 
weakens the plates and tubes. 

Table 59 gives the results of a number of tests made on loco- 
motive boiler tubes with different thicknesses and characters of 
scale. The diversity of the results indicates the futility of bas- 
ing the decrease in conductivity on the thickness of the scale. For 
example, test No. 1 shows a decrease in conductivity of 9.1 per 
cent for a scale .02 inch thick, while No. 16 shows a decrease 
of only 6.75 per cent for a scale over 6.5 times as thick. The 
scale in each case was even, hard, and dense. Again, No. 8 with 
a very soft scale .042 inch thick gives a decrease in conductivity 
of 9.54 per cent, whereas No. 14, also very soft but twice as thick, gives 
a decrease of only 4.95 per cent. No doubt the heat transmission is a 
function of the chemical as well as the physical properties, but 
further experiments are necessary before any specific conclusion can be 
drawn. 

Waters containing acids, organic matter, and magnesium chloride 
and sulphate tend to corrode the boiler, and those containing sodium 
carbonate, organic matter, and alkalies induce priming. Even distilled 
water, as obtained from a surface condenser, is a solvent of iron to a 
certain extent and causes corrosion and pitting. Table 60 gives some 
idea of the character and extent of impurities in water from various 
localities, with an analysis of the scale produced by the water and the 
trouble in the boiler arising from its use. 

It is impossible to judge the quality of feed water merely by the 
grains of solids per gallon, since a large amount of soluble salt such as 
sodium chloride will not be as deleterious as a very small amount of 
calcium sulphate. 

471 



172 



STEAM POWER PLANT ENGINEERING 



TABLE 59. 

INFLUENCE OF SCALE ON HEAT TRANSMISSION. 

(Locomotive Boiler Tubes.) 



No. 


Thickness of Scale, 
Inches. 


Character of Scale. 


Decrease in Con- 
ductivity due to 
Scale. Per cent. 


1 


.02 

.02 

.033 

.033 

.038 

.04 

.04 

.042 

.047 

.065 

.07 

.07 

.085 

.089 

.11 

.13 


Hard, dense 

Hard 

Soft 

Very hard 

Medium 

Soft, porous 

Hard, dense 

Very soft 

Hard 

Medium 

Soft 

Hard 

Soft, porous 

Very soft 

Hard, porous 

Hard, dense 


9 1 


2 


2 02 


3 


4 3 


4 


3.5 


5 


4.03 


6 


6.82 


7 


3.07 


8 


9.54 


9 


2 75 


10 


2.39 


11 


2.38 


12 


4.43 


13 


19.0 


14 


4.95 


15 


16.73 


16 


6.75 







From tests conducted at the University of Illinois, Railroad Gazette, Jan. 27, 1899, June 14, 
1901. See also Engineering Record, Jan. 14, 1905, p. 53; Power, February, 1903, p. 70; 
Street Railway Review, July 15, 1901, p. 415. 

The following is a rough rating according to the number of grains of 
incrusting solids per United States gallon: 

Less than 

8 grains very good. 

12 to 15 grains good. 

15 to 20 grains fair. 

20 to 30 grains bad. 

Over 30 grains very bad. 

This applies to calcium carbonate, magnesium carbonate, and mag- 
nesium chloride. For water containing sulphate of calcium and mag- 
nesium, divide the first column by 4 for the same rating. 

On account of the great variety of possible impurities the proper 
treatment to be adopted can be determined only by chemical analysis 
of the feed water in each case. 

Table 61, compiled by the Hartford Steam Boiler Inspection and 
Insurance Company, shows the number of boilers inspected by that 
company during the year 1907 and the number found defective from 
various causes. 



FEED-WATER PURIFIERS AND HEATERS 



473 



m 

GO 
>H 

< 
<! 
H 

d 3 

pq d 
< o 

H « 
Q 

H 
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474 



STEAM POWER PLANT ENGINEERING 



TABLE 61. 

SUMMARY OF INSPECTORS' REPORTS FOR THE YEAR 1907. 
(Hartford Steam Boiler Inspection and Insurance Company.) 



Nature of Defects. 



Whole Number. 


Dangerous. 


18,917 


1,315 


38,427 


1,333 


3,010 


258 


12,802 


528 


10,230 


768 


2,219 


578 


6,363 


699 


7,564 


396 


3,551 


568 


4,878 


499 


898 


92 


3,582 


823 


1,764 


238 


11,357 


1,599 


8,266 


3,054 


1.947 


563 


5,557 


430 


3,008 


707 


4,216 


1,250 


413 


156 


1,231 


415 


1,211 


407 


7,651 


465 


194 


194 


27 


10 



Cases of deposit of sediment. . . 
Cases of incrustation and scale 

Cases of internal grooving 

Cases of internal corrosion. . . . 
Cases of external corrosion. . . . 
Defective braces and stays 

Settings defective 

Furnaces out of shape 

Fractured plates 

Burned plates 

Laminated plates 

Cases of defective riveting. . . . 

Defective heads 

Leakage around tubes 

Cases of defective tubes 

Tubes too light 

Leakage at joints 

Water gauges defective 

Blow-offs defective 

Cases of deficiency of water . . . 

Safety valves overloaded 

Safety valves defective 

Pressure gauges defective 

Without pressure gauges 

Unclassified defects 

Totals 



159,283 



17,345 



The neutralization or elimination of the impurities may be effected 
by one of the following methods: 

1. Chemically. 

Boiler compounds. 
Purifying plants. 

2. Mechanically. 

Filters. 
Blow-off. 
Tube cleaners. 

3. Thermally. 

Feed-water heater. 
Distillation. 









FEED-WATER PURIFIERS AND HEATERS 



475 



The following chart (" Boiler Waters," W. W. Christie) outlines some 
of the troubles arising from feed water, their cause and means for 
preventing them. 



Trouble. 



Cause. 



Remedy or Palliation. 



Incrustation. 



Corrosion. 



Priming . 



Sediment, mud, clay, etc.. 
Readily soluble salts 

Bicarbonate of magnesia, 
lime, iron 

Organic matter. . 

Sulphate of lime < 

Organic matter 

Grease 

Chloride or sulphate of ) 

magnesium ( 

Sugar } 

Acid f 

Dissolved carbonic acid and 
oxygen 

Electrolytic action 

Sewage ■] 

Alkalies 

Carbonate of soda in large ) 
quantities ) 



Filtration. 

Blowing off. 

Blowing off. 

Heating feed and precipitate. 

Caustic soda. 

Lime. 

Magnesia. 

See below. 

Sodium carbonate. 

Barium chloride. 

Precipitate with alum ) 

Precipitate with ferric > and filter 

chloride ) 

Slaked lime 



Carbonate of soda 
Carbonate of soda. 



and filter 



Alkali. 

Slaked lime. 

Caustic soda. 

Heating. 

Zinc plates. 

Precipitation with alum or ferric 

chloride and filter. 
Heating feed and precipitate. 

Barium chloride. 



Analysis of Water for Softening by Chemical Processes: Jour. Soc. Chem. Ind., 
May, 1899, p. 520. Volumetric Determination of Calcium and Magnesium in Water: 
Jour. Soc. Chem. Ind., May 31, 1901, p. 507. Simple Tests for Boiler Water: Soc. 
Engng., May, 1904, p. 238. Qualitative Tests of Feed Water: Power, Dec, 1904, 
p. 756. New Testing Apparatus for Boiler Feed Water: West. Elecn., Aug. 1, 1903, 
p. 85. A Simple Method of Calculating Water Analyses and Amounts of Substances 
to be Added for Preventing Scale and Corrosion: Jour. Frank. Inst., Vol. CLIX, p. 217. 
Description of Dearborn Drug and Chemical Co.'s Laboratories for Analyses of Boiler 
Feed Water: St. Ry. Review, Sept. 15, 1901. 

Boiler Corrosion: Power, Jan., 1906, Oct., 1905, p. 591, Dec, 1900; Eng. Rev., 
Oct., 1904, p. 12; Eng. Mag., Dec, 1905, p. 425, Oct., 1900, p. 118; Engng., Oct. 10, 
1902, p. 482; Engr. U.S., Jan. 1, 1907, p. 103, May 1, 1902, p. 388; Engr., Lond., 
Aug. 5, 1904, p. 131, July 29, 1904, p. 115, Dec. 4, 1896, p. 574; Elec Age, Dec, 
1905, p. 456; Am. Elecn., Aug., 1905, p. 436, April, 1902, p. 184; Jour. Am. Soc 
Nav. Engrs., May, 1902; Mines and Min., Sept., 1903; Stahl u. Eisen, Jan. 15, 1904. 

Boiler Incrustation: Am. Elecn., April, 1904, p. 206, Dec, 1901, p. 576, May, 
1901, p. 220, Oct., 1898, p. 473; Cassier's Mag., July, 1903, p. 273; Chem. News, 



476 STEAM POWER PLANT ENGINEERING 

Oct. 18, 1901, p. 191; Engr., Lond., Jan. 21, 1898, p. 52; Engr. U.S., April 16, 1906, 
March 15, 1904, p. 202, May 15, 1904, p. 354, Sept. 1, 1904, p. 608; Ice and Refrig., 
Nov., 1905, p. 173. Chemistry of Scale: Jour. Frank. Inst., Aug., 1901, p. 113, Aug. 
1891, p. 145. Boiler Scale: Power, Dec, 1905, p. 779 ; St. Ry. Review, April 2, 1904, 
p. 545. Scale Prevention: Am. Elecn., Dec, 1901, p. 578; Am. Engr., May, 1900, 
p. 138; Eng. Mag., 1897, 12-959, 13^74, 232, 419; Elec Engr., Lond., July 20, 1900, 
p. 91; Engrs. Gaz., July, 1902. 

Foaming. — Foaming Water and Scaling Water for Locomotive Boilers: Eng. 
News, July 21, 1904, p. 71, Sept. 1, 1904, p. 198; Foaming and Priming: Soc Nav. 
Arch, and Marine Engrs., 1902; R.R. Gazette, Oct. 12, 1900; Christie, Boiler Waters, 
Chap. V; Stromeyer, Steam Boilers, pp. 67-83; Rowan, Modern Steam Boilers, 
pp. 321-354. 

239. Chemical Purification. — Chemical treatment of boiler feed 
water has been remarkably developed during recent years and a 
number of manufacturing concerns make this their sole business. The 
two most common systems of chemical treatment involve (1) boiler 
compounds and (2) purifying plants. In the former the necessary 
chemical action takes place inside the boiler and in the latter the 
water is purified before it enters the boiler. In either case the 
usual procedure is to submit for analysis a sample of the feed water 
and the resulting scale to a competent chemist who will specify the 
character and quantity of chemicals necessary to bring about the 
desired result. 

240. Boiler Compound. — The object of treatment with boiler com- 
pounds is to neutralize the evil effects of the impurities in the feed 
water or to change them into others which are less objectionable and 
which are easily removed. When properly compounded and intro- 
duced into the boiler such preparations are of great benefit and prac- 
tically overcome the deleterious effects, but when improperly used 
they may produce even greater troubles than the impurities which they 
are expected to eliminate. 

Boiler compounds may be divided into three classes : 

1. Those converting the scale-forming elements into new sub- 
stances which will not form a hard, resisting scale and which are 
readily removed by skimming, blowing off, or by tube cleaners. For 
example, feed water containing sulphates of lime and magnesia will 
form a dense, tenacious scale. If carbonate of soda be added in correct 
amount, the sulphates are converted into insoluble carbonates which 
are precipitated and form scale varying from a more or less porous, 
friable crust to a soft " mush " or mud. The resulting sulphate of 
soda remains in solution and does not form scale unless allowed to 
concentrate, and this is prevented by blowing off. An excess of soda 



FEED-WATER PURIFIERS AND HEATERS 477 

is apt to cause foaming and at high temperatures is liable to attack 
the inside of gauge glasses. Bisodium and trisodium phosphate, 
sodium t annate, fluoride of sodium, sugar, etc., have all proved satis- 
factory, but as each case requires special treatment no detailed dis- 
cussion is possible within the scope of this work and the reader is 
referred to the accompanying bibliography. 

2. Those enveloping the newly precipitated scale-forming crystals 
with a surface which prevents them from cementing together. The 
ingredients used to bring about this result are starches, woody fibers, 
dextrine, slippery elm, and the like. 

3. Those preventing the formation of hard scale by a solvent or 
" rotting " action, as kerosene and petroleum oils. 

Boiler Compounds. — Use of Compounds: Eng. News, July 27, 1905, p. 112; 
Am. Mach., Dec. 7, 1899, p. 115, Oct. 26, 1899, p. 1014; Power, Aug., 1903; Eng. 
and Min. Jour., Aug. 12, 1905, p. 253. 

241. Use of Kerosene and Petroleum Oils in Boiler Feed Water. — 

Kerosene oil and other refined petroleum oils are sometimes used with 
good effect in boilers to prevent scale from adhering. These oils are 
said to change the deposit of lime from a hard scale to a friable material 
which may be easily removed. They are ordinarily fed to the boiler 
with the feed water, drop by drop, through a sight feed apparatus 
similar to a cylinder oil lubricator. From extended experiments 
made on a 100-horse-power tubular boiler fed with water containing 6.5 
grains of solid matter per gallon it was found that one quart of kero- 
sene per day was sufficient to keep the boiler entirely free from scale. 
Prior to the introduction of the oil the water had a corrosive action 
upon some of the fittings attached to the boiler, but after the oil had 
been used for a few months it was found that the corrosive action had 
ceased. In another case 40 gallons of kerosene were used in 24 hours 
in a steamer of about 3000 horse power. These boilers showed no 
incrustation but considerable corrosion. Evidently oil does not have 
the same effect or give the desired results in all cases. Kerosene used 
in moderate quantities will not cause foaming. Crude oil should never 
be used, as the heavy residue causes the formation of a tough, imper- 
vious scale productive of bagged sheets and collapsed flues. 

Use of Kerosene in Boilers: Engr. U.S., Sept. 15, 1905, p. 634; Eng. News, May 
24, 1890, p. 497; Power, Aug., 1895, p. 13, May, 1896, p. 16; Trans. A.S.M.E., 9-247, 
11-937; Locomotive, July, 1890, p. 97. 

242. Use of Zinc in Boilers. — Zinc is often introduced into boilers 
to prevent corrosion. The theory is that a feeble but continuous cur- 



478 STEAM POWER PLANT ENGINEERING 

rent of hydrogen is generated over the whole extent of the iron by 
electrolytic action. The bubbles of hydrogen formed isolate the 
metallic surface from scale-forming substances. If there is but a little 
of the scale-forming element it is precipitated and reduced to mud; if 
there is considerable, coherent scale is produced which takes the form 
of the iron surface but does not adhere to it, being prevented from doing 
so by the intervening bubbles of hydrogen. Zinc is ordinarily sus- 
pended in the water space of the boiler in the shape of blocks, slabs, 
or as shavings in a perforated vessel. Electrical connection between 
the metallic surfaces is essential. Rolled zinc slabs 12x6xJ inches 
have found much favor in marine practice. Generally speaking one 
square inch of zinc surface is sufficient for every 50 pounds of water 
in the boiler, though the quantity placed in the boiler should vary with 
the hardness. The British Admiralty recommends the renewing of the 
zinc slabs whenever the decay has penetrated to a depth of J inch 
below the surface. Zinc does not prevent corrosion or scale formation 
in all cases and may even aggravate the trouble. 

Use of Zinc in Boilers: Am. Elecn., Dec, 1901, p. 572; Kent, Steam Boiler Econ- 
omy, p. 318; Christie, Boiler Waters, p. 137; Stromeyer, Marine Boiler Management 
and Construction, p. 81. 

243. Methods of Introducing Compounds. — Boiler compounds may 
be introduced into the boiler continuously or intermittently. Small 
quantities introduced continuously or at short intervals are more 
effective than large quantities at long intervals. Continuous feeding 
is ordinarily brought about by connecting the suction side of the feed 
pump with a reservoir containing the compound in solution, arranged 
similarly to an ordinary cylinder oil lubricator. In large plants an 
independent pump is often used to force the solution into the feed line. 
Intermittent feeding is brought about by temporarily connecting the 
suction of the feed pump with the reservoir containing the compound. 
The use of boiler compounds does not necessarily prevent scale from 
forming in time, though it will reduce the evil to a minimum. In some 
instances where compounds are used it is found necessary to run a 
tube cleaner through the tubes at certain intervals, in others such a 
course has not been found necessary. 

244. Weight of Compound Required. — The weight of compound 
introduced depends upon the nature of the reagents used and the 
character of the feed water, and ranges from a few ounces to several 
pounds per 100 gallons of feed water. For example, water containing 
4 grains of calcium sulphate and 6 grains of magnesium sulphate per 
gallon will require 3.57 pounds of carbonate of soda per 1000 gallons 



FEED-WATER PURIFIERS AND HEATERS 479 

of water for the reduction of the sulphates. The chemical reaction 
and analysis is as follows: 

CaS0 4 + Na 2 C0 3 = CaC0 3 + Na 2 S0 4 
MgS0 4 + Na 2 C0 3 = MgC0 3 + Na 2 S0 4 

If x = grains of Na 2 C0 3 necessary for the calcium, 

CaS0 4 : Na 2 C0 3 + 10H 2 O = 4 : x. 
40 + 32 + 4 X 16 : 2 X 23 + 12 + 3 X 16 + 10 (2 + 16) = 4 : x. 

x = 8.41 grains. 

If y = grains of Na 2 C0 3 necessary for the magnesium, 

MgS0 4 : Na 2 C0 3 + 10H 2 O = 6 : y. 
24 + 32 + 4 X 16 : 2 X 23 + 12 + 3 X 16 + 10 (2 + 16) = 6 : y. 

y = 14.3. 

The total weight of carbonate of soda per 1000 gallons is therefore 

1000 (14.3 + 8.41) = 22,710 grains. 
= 3.24 pounds. 

This amount would effect the desired result if the chemical reaction 
is permitted to take place for some time, otherwise an excess of reagent 
is necessary. 

245. Mechanical Purification. — Waters containing sand, mud, 
organic matter, and in fact all matter which is not in solution or in 
chemical combination with the water may be purified by mechanical 
filtration. Mud and sand may be eliminated by simply permitting the 
water to stand for some time in settling tanks. Suspended matter 
which will not gravitate to the bottom may be removed by filtering the 
water through coke, cloth, excelsior, or the like. Filters should be in 
duplicate for continuity of operation. 

Vegetable and other organic impurities commonly float on the sur- 
face of the water when the boiler is making steam, and may be blown 
out through a " surface blow-out." (See paragraph 82.) 

Precipitated matter may be ejected from the boiler by fre- 
quent blowing off before it has time to adhere and bake to a crust. 
This procedure is particularly essential when boiler compounds are 
used. 

For description and use of mechanically operated tube cleaner see 
paragraph 86. 

246. Thermal Purification. — (See also Live Steam Purifiers, para- 
graph 263.) The carbonates of lime and magnesia are held in solution 



480 STEAM POWER PLANT ENGINEERING 

in fresh water by an excess of carbon dioxide and are completely pre- 
cipitated by boiling. At ordinary temperatures carbonate of lime is 
soluble in approximately 20,000 times its volume of water, at 212 
degrees F. it is slightly soluble, and at 290 degrees it is insoluble. Sul- 
phate of lime is much more soluble in cold than in hot water, and is com- 
pletely precipitated at 290 degrees. (Revue de Mecanique, November, 
1901, pp. 508, 743.) Thus it will be seen that a feed heater may be 
relied upon to remove part or all of the lime, depending upon the tem- 
perature to which the water is raised and the time in which the pre- 
cipitation is permitted to take place. 

Influence of Temperature and Concentration on the Saline Constituents of Boiler 
Water : Jour. Soc. Chem. Ind., Oct. 31, 1900, p. 885. Solubility of Sulphate of Lime: 
Rev. de Mecanique, Jan., 1901, p. 5, Nov., 1901, p. 508. 

247. Purifying Plants. — The function of a purifying plant is the 
elimination of all impurities from the feed water before it enters the 
boiler. In the Scaife system for water purification feed water first 
enters the heater, where it attains a temperature of from 200 to 210 
degrees F. As a portion of the free C0 2 is driven off by the heat the 
carbonates of lime and magnesia are precipitated and are deposited in 
removable pans inside the heater. On its way the heated water is 
forced by the boiler feed pump into a large precipitating tank, where 
the necessary chemicals are introduced by means of two small pumps. 
These pumps take the solution of chemicals from the solution tanks which 
hold a sufficient quantity to operate the plant from eight to twelve hours. 
The precipitating tank is so constructed as to cause intimate and thorough 
mixing of the chemicals with the water. Thus the acids are neutralized, 
and the scale-forming substances are precipitated by being changed to 
insoluble substances which sink to the bottom of the precipitating tank 
whence they are readily removed. Some of the lighter substances 
remaining in suspension are carried along with the water as it passes 
into the filters, which effectively remove all suspended matter. This 
system is continuous in operation, and purification is accomplished 
without appreciably retarding the onward flow of feed water. Fig. 237 
shows a modification of the system. The chemicals are pumped from 
the " chemical tank "into the " solution tanks," where the feed water 
and chemical solution are thoroughly mixed. The treated water is taken 
from these tanks and pumped into the " precipitating tanks " where a 
large portion of the scale-forming element is precipitated. From the 
precipitating tanks the water is forced through a series of filters to the 
boiler. 

Fig. 238 illustrates the We-Fu-Go system of water purification. In 



FEED-WATER PURIFIERS AND HEATERS 



481 



PRECIPITATING 
TANKS 




L_JL_Ji J i i i i i ! L_j L_J L.i 

Fig. 237. General Arrangement of Scaife System of Feed-Water Purification. 




Fig. 238. General Arrangement of We-Fu-Go System of Feed- Water Purification. 



482 



STEAM POWER PLANT ENGINEERING 



this installation the water supply first enters the settling or treating 
tanks into which the chemicals are fed. A thorough mixture is effected 
by the use of the two armed paddles located near the bottom of the 
tanks. From the treating tanks the water flows by gravity into the 
filters, which remove all remaining impure solid matter which does not 
settle to the bottom of the treating tank. The pipes conducting the 
water from the settling tanks to the filter are fitted with a flexible joint 
and float so that the outlets are near the surface at all times, rising and 
falling with the water level. From the filters the purified water gravi- 
tates into the clear water storage reservoir, from which it is pumped 
into an open heater and thence to the boiler. This system is intermit- 
tent in operation, and in order to provide sufficient time for thorough 
chemical treatment of large quantities, two or more settling tanks are 
employed. Both the We-Fu-Go and Scaife systems are modified in a 
number of ways to meet different conditions. 




Fig. 238a. Anderson System for Preventing Corrosion in Condensers. 

Fig. 238a shows the general arrangement of the Anderson system 
for preventing corrosion in condensers and removing oil from condensed 
steam. The method consists in injecting into the exhaust steam as it 
passes from the preheater to the condenser a solution containing a 
coagulant which changes the emulsion of the cylinder oil to a flaky 
condition so that it may be separated by settling, flotation, or filtering. 
The air pump delivers the water to the settling tank F, whence it is 
taken to the open gravity filters G, G, of a superficial area proportional 
to the amount of water to be passed and containing a filter bed of four 
feet of crushed quartz. This will run about four days without any 
marked difference in efficiency, after which time the bed is stirred to a 
depth of two feet by mechanical agitators and flushed with clean water, 
by which all impurities are carried to the sewer. The solution is pre- 



FEED-WATER PURIFIERS AND HEATERS 483 

pared in tank A, in which the water level is preserved by a ball float 
and into which filtered water is admitted through pipe B, while the 
substance with which the water is treated is pumped in through the 
pipe D by a small pump operated from the main engine. The flow to 
the " rose head " above the condenser is controlled by the valve E, 
and a meter in this pipe records the amount being fed. The water 
ordinarily required for " make up " is sufficient to carry in the solution. 
There is very little loss of water, and the rapid corrosion of the con- 
denser tubes, which has been so great an obstacle to the successful use 
of surface condensers, is much reduced. The chemicals used perform 
a twofold duty, viz., to neutralize the water and make it chemically 
inactive and to coagulate the oily matter contained in the steam so that 
mechanical filtration is possible. (Power, June 1903, p. 304.). 

Water-softening plants cost from $4 to $5 per horse power for 
plants of 1000 horse power and less, from $3 to $4 for plants of 
1000 to 2000 horse power, and as low as $1.50 for plants of 5000 
horse power or more. The depreciation of wooden tanks is as high as 
15 per cent a year, while that of steel tanks should not be greater than 
5 per cent. Unless wooden tanks are considerably cheaper than steel 
tanks they are not a good investment. The cost of water purification 
varies from a fraction of a cent to 2 cents per 1000 gallons, depending 
upon the size of the plant and the quantity and character of the impuri- 
ties. (American Electrician, March, 1905, p. 125.) 

Feed Water Purification: Am. Elecn., March, 1900, p. 145, April, 1900, p. 190, 
Dec, 1904, p. 618; Cassier's Mag., April, 1904, p. 506; Engng., Oct. 25, 1901, p. 595; 
Eng. News, May 22, 1902, p. 408; Eng. Rec, April 5, 1902, p. 322, June 10, 1905; 
Jour. Amer. Chem. Soc, Nov., 1893, p. 610; Jour. Soc. Chem. Ind., Aug., 1901, 
p. 828; Power, Nov., 1900, p. 7, Sept., 1902, p. 33, Nov., 1904, p. 693; R.R. Gaz., Aug. 
24, 1900, p. 568; Elec. Rev., Nov. 12, 1904; Engr. U.S., Oct. 15, 1903, Jan. 1, 1906; 
Elec. World, Sept. 1, 1906. 

Water-Softening Plants. — Water-Softening Processes: Prac. Engr. U.S., Mar , 1910. 
Four Systems of Softening Water for Industrial Purposes: Eng. News, July 2, 1903, 
p. 4. An Inquiry into the Working of Various Water Softeners: Inst, of Mech. Engrs., 
Dec, 1903. Report on Soft Water for Locomotive Plants: Eng. News, March 17, 
1904. The Development of Water Purification in the U.S.: R.R. Gaz., 38-19. Gen- 
eral Information on Water Softening: Eng. News, May 26, 1904, p. 500, 508, June 
2, 1904, p. 530; Eng. Rec, Oct. 24, 1903, p. 483; Loco. Engng., Nov., 1903, p. 501; 
Elec. Engr., Lond., April 21, 1905; Ir. and Coal Tr. Rev., Sept. 1, 1905; Jour. W. Soc. 
Engrs., Dec, 1905; Eng. and Min. Jour., Dec. 2, 1905; Am. Engr. and R.R. Jour., 
Jan., 1905. 

Harris- Anderson Water Softener: Engng., Aug. 15, 1902, July 10, 1903. 

Kennicott Water Softeners: Am. Elecn., Nov., 1902, p. 545; Eng. News, May 15, 
1902, p. 386; Eng. Rec, May 3, 1902, p. 419; St. Ry. Jour., April 2, 1904, 
p. 545. 



484 STEAM POWER PLANT ENGINEERING 

Burt Continuous Water-Softening Process: Eng. News, Sept., 15, 1904, p. 238; 
Engr. U.S., July 15, 1905, p. 426. 

Bachman Method of Water Purification: St. Ry. Review, May 15, 1900, p. 282. 

We-Fu-Go and Scaife Systems: St. Ry. Rev., Oct., 1901, p. 771; Engr. U.S., 
Jan. 1, 1903, p. 90. 

Holmes System of Water Purification: Power, April, 1905, p. 248. 

The American Water Purifier and Softener: Eng. U.S., Aug. 1, 1904, p. 551. 

248. Economy of Preheating Feed Water. — Although a feed water 
heater acts to some extent as a purifier its primary function is that of 
heating the feed water. Generally speaking, for every 10 degrees 
that the feed water is heated there is a gain in heat of 1 per cent and a 
corresponding saving of coal, if the heat which warms the feed water 
would otherwise be wasted. Again, the smaller the difference in 
temperature between the steam and the feed water the less will be the 
strain on the boiler shell due to unequal expansion and contraction, an 
item of no small consequence. 

If X represents the total heat of one pound of steam above 32 
degrees F., t the temperature of the cold water, and t the temperature 
of the water leaving the heater, then S, the per cent gain in heat due 
to heating the feed water, may be expressed 

S = 100 x ( *~* o) ^ • ( 104 ) 

X~(t -32) 

The expression is not theoretically correct, since it assumes a con- 
stant value of unity for the specific heat, whereas the specific heat 
varies with the temperature. The variation is so slight, however, 
that it may be neglected for all practical purposes. 

Example: Steam pressure 100 pounds gauge; temperature of water 
entering heater 80 degrees F.; temperature of water leaving heater 
210 degrees F. Required, saving due to heating the feed water. 

Here \ (from steam tables) is 1185, t = 80, t = 210. 

S =10 o (210-80) 



1185- (80-32) 
= 11.42 per cent. 



This formula gives the thermal saving only, and the first cost of the 
heater, interest, depreciation, attendance, and repairs must be taken 
into consideration before the net saving measured in dollars and cents 
is ascertained. In the average installation the net saving is a sub- 
stantial one. 



FEED-WATER PURIFIERS AND HEATERS 



485 



Table 62 based upon formula (104) may be used in determining the 
percentages of saving due to the increase in feed-water temperature. 

TABLE 62. 

PERCENTAGE OF SAVING FOR EACH DEGREE OF INCREASE IN TEMPERATURE 

OF FEED WATER. 



Initial 








Boiler Pressure above Atmosphere. 








Temp. 
























of Feed. 





20 


40 


60 


80 


100 


120 


140 


160 


180 


200 


32 


.0872 


.0861 


.0855 


.0851 


.0847 


.0844 


.0841 


.0839 


.0837 


.0835 


.0833 


40 


.0878 


.0867 


.0861 


.0856 


.0853 


.0850 


.0847 


.0845 


.0843 


.0841 


.0839 


50 


.0886 


.0875 


.0868 


.0864 


.0860 


.0857 


.0854 


.0852 


.0850 


.0848 


.0846 


60 


.0894 


.0883 


.0876 


.0872 


.0867 


.0864 


.0862 


.0859 


.0856 


.0855 


.0853 


70 


.0902 


.0890 


.0884 


.0879 


.0875 


.0872 


.0869 


.0867 


.0864 


.0862 


.0860 


80 


.0910 


.0898 


.0891 


.0887 


.0883 


.0879 


.0877 


.0874 


.0872 


.0870 


.0868 


90 


.0919 


.0907 


.0900 


.0895 


.0888 


.0887 


.0884 


.0883 


.0879 


.0877 


.0875 


100 


.0927 


.0915 


.0908 


.0903 


.0899 


.0895 


.0892 


.0890 


.0887 


.0885 


.0883 


110 


.0936 


.0923 


.0916 


.0911 


.0907 


.0903 


.0900 


.0898 


.0895 


.0893 


.0891 


120 


.0945 


.0932 


.0925 


.0919 


.0915 


.0911 


.0908 


.0906 


.0903 


.0901 


.0899 


130 


.0954 


.0941 


.0934 


.0928 


.0924 


.0920 


.0917 


.0914 


.0912 


.0909 


.0907 


140 


.0963 


.0950 


.0943 


.0937 


.0932 


.0929 


.0925 


.0923 


.0920 


.0918 


.0916 


150 


.0973 


.0959 


.0951 


.0946 


.0941 


.0937 


.0934 


.0931 


.0929 


.0926 


.0924 


160 


.0982 


.0968 


.0961 


.0955 


.0950 


.0946 


.0943 


.0940 


.0937 


.0935 


.0933 


170 


.0992 


.0978 


.0970 


.0964 


.0959 


.0955 


.0952 


.0949 


.0946 


.0944 


.0941 


180 


.1002 


.0988 


.0981 


.0973 


.0969 


.0965 


.0961 


.0958 


.0955 


.0953 


.0951 


190 


.1012 


.0998 


.0989 


.0983 


.0978 


.0974 


.0971 


.0968 


.0964 


.0962 


.0960 


200 


.1022 


.1008 


.0999 


.0993 


.0988 


.0984 


.0980 


.0977 


.0974 


.0972 


.0969 


210 


.1033 


.1018 


.1009 


.1003 


.0998 


.0994 


.0990 


.0987 


.0984 


.0981 


.0979 


220 




.1029 


.1019 


.1013 


.1008 


.1004 


.1000 


.0997 


.0994 


.0991 


.0989 


230 




.1039 


.1031 


.1024 


.1018 


.1012 


.1010 


.1007 


.1003 


.1001 


.0999 


240 




.1050 


.1041 


.1034 


.1029 


.1024 


.1020 


.1017 


.1014 


.1011 


.1009 


250 




.1062 


.1052 


.1045 


.1040 


.1035 


.1031 


.1027 


.1025 


.1022 


.1019 



Multiply the factor in the table corresponding to any given initial temperature of feed water 
and boiler pressure by the total rise in feed-water temperature; the product will be the percent- 
age of saving. 

Feed Water Heating. — How Should Feed Water be Heated ? — Power, July, 1907, 
p. 456; Feed Water Heating: Engr. U.S., Jan. 1, 1906, p. 8, Aug. 15, 1904, p. 15; 
St. Ry. Jour., July 22, 1905, p. 145; Am. Elecn., Dec, 1904, p. 570; Am. Elecn., 
Nov., 1904; Engr., Lond., July 28, 1905. 

249. Classification of Feed- Water Heaters. — Feed-water heaters 
may be classified according to the source of heat, as 

1. Exhaust steam, in which the heat is received from the exhaust of 
engines, pumps, etc. 

2. Flue gas, in which the waste chimney gases are the source of the 
heat. 

3. Live steam purifiers, or those using steam at boiler pressures; or 
according to the method of heat transmission, as 



486 STEAM POWER PLANT ENGINEERING 

1. Open heaters, in which the steam and feed water mingle and the 
steam in condensing gives up its heat directly to the water. 

2. Closed heaters, in which the steam and water are in separate 
chambers and the steam gives up its heat to the water by conduction. 

Heaters may also be classified according to the pressure of the heat- 
ing steam, as 

1. Vacuum or primary, in which the pressure is less than atmos- 
pheric and applies particularly to heaters utilizing the exhaust of con- 
densing engines. These are always of the closed type. Open heaters 
in which the pressure is less than atmospheric are not usually classed 
as vacuum heaters. 

2. Atmospheric or secondary, in which the pressure is atmospheric 
or, literally, that corresponding to the back pressure on the engines 
and pumps. 

3. Pressure, in which the pressure corresponds to that in the boiler 
and in which the heat is used primarily for purifying purposes. 

Heaters may be still further classified as 

1. Induced, in which only such steam is admitted as is induced by 
its condensation. That is, the feed water condenses the steam. This 
creates a partial vacuum which draws in more steam. 

2. Through, in which all the steam is forced through the heater 
irrespective of condensation. 

CLASSIFICATION OF A FEW TYPICAL HEATERS. 



Exhaust steam 



Open Atmospheric 

~, . ( Atmospheric 

Closed.. ]„ 

( Vacuum or pressure 



Cochrane 

Hoppes 

Stillwell 

Webster 

Wainwright ) Water 

Wheeler . . . ) Tube 

Otis ) Steam 

Berryman . j Tube 

{Green 
American 
Sturtevant 

Live Steam Open Pressure ] ~ ,, 

^ ( Baragwanath 

250. Open Heaters. — Fig. 239 gives a sectional view of a Cochrane 
special feed heater and receiver and is a typical example of an open 
heater. Exhaust steam enters the heater through a fluted oil separa- 
tor as indicated, and passes out at the top, while the oily drips are 
automatically drained to waste by a suitable ventilated float. The 
feed water enters through an automatic valve and is distributed over 



FEED-WATER PURIFIERS AND HEATERS 



487 



a series of copper trays so arranged and constructed that the water is 
forced to fall in a finely divided stream before reaching the reservoir in 
the bottom. The steam coming in contact with the water particles 
gives up latent heat and condenses. Much of the scale-forming ele- 
ment is deposited on the surface of the trays, from which it is readily 
removed. The suspended matter is eliminated by a coke filter in the 




Fig. 239. Cochrane Special Heater and Receiver. 

bottom of the chamber, and the floating impurities are decanted by a 
skimmer or overflow weir. The particular heater shown in the illustra- 
tion is especially designed for use in a steam-heating plant; i.e., 
besides performing all the functions of an open heater, it provides for 
the reception and heating of the condensation returned to it from the 
heating system. 



488 



STEAM POWER PLANT ENGINEERING 



Fig. 240 gives a sectional view of a Webster " star vacuum " heater. 
W^ater enters the heater through balanced valve F, which is controlled 
by float E, and is deflected over a series of perforated copper trays T, T. 
Exhaust steam enters at A, passes through oil filter S, and, mingling 
with the finely divided streams of water, gives up its latent heat and is 
condensed. Only so much steam enters the heater as is condensed by 
the feed water. The condensed steam and feed water fall to the bottom 




Fig. 240. Section Through Webster Heater. 



of the upper chamber, maintaining a practically constant level WW. 
From this upper or heater chamber the water gravitates to the settling 
chamber at the bottom, through down-cast pipe CB. From the set- 
tling chamber the water rises through perforated screen M and filtering 
material P to the outlet 0. A large portion of the scale-forming ele- 
ment is precipitated on the trays or collects in the settling chamber at 
the bottom. 






FEED-WATER PURIFIERS AND HEATERS 



489 



Fig. 241 shows a section through a Hoppes open heater, illustrating 
the " pan " type. Exhaust steam enters at H, passes through oil filter 0, 
and completely surrounds pans T, T. The feed water enters at B, 
and the rate of flow is regulated by valve F, which is controlled by a 




Fig. 241. Hoppes Horizontal Feed-Water Heater. 

suitable float in the lower part of the chamber. The water in flowing 
over the sides and bottoms of the pans comes in direct contact with 
the steam. 

251. Combined Open Heater and Chemical Purifier. — Combined feed- 
water heaters and chemical purifiers are finding increased favor with 
engineers in many districts where the feed water is particularly bad. 
A description of the Webster combination will be found in Part II of 
the general catalogue issued by the Warren Webster Company, Camden, 
N.J. A description of the Cochrane-Sorge combined heater and chem- 
ical purifier will be found in the heater catalogue issued by the Harrison 
Safety Boiler Works, Philadelphia, Pa. 

252, Temperatures in Open Heaters. — The temperature to which 
feed water is raised in an open heater may be determined as follows : 

Let A represent the total heat of steam corresponding to the pressure 
in the heater, 
t the temperature of the water entering heater, 
t the temperature of the water leaving heater, and 
S the ratio of exhaust steam to the feed water, by weight. 



490 



STEAM POWER PLANT ENGINEERING 



Then, allowing a loss of 10 per cent due to radiation, etc., 
0.9 S (a — t + 32) will be the B.T.U. given up by the exhaust steam to 
each pound of feed water, and (t — 1 ) will be the B.T.U. absorbed by 
each pound of water. 

Therefore 0.9 S (X - t + 32) = t 



t , from which 



= t + 0.9 S (X + 32) 
1 + 0.9 S 



(105) 



If more steam passes through the heater than can be condensed by 
the feed water, then this equation gives t a fictitious value; in other 
words, t can never be greater than the temperature of the exhaust 
steam. 

Substituting t = 212, the maximum obtainable temperature with 
exhaust steam at atmospheric pressure, and solving for S, we find that 
only 17 per cent of the main engine exhaust is necessary to heat the 
feed water to a maximum. t is assumed to be 60 degrees F. 

Table 63 has been determined from this equation and gives the final 
temperatures obtainable in open heaters for various conditions of 
operation. 

TABLE 63. 





FINAL 


FEED- 


WATER 


TEMPERATURES. 


OPEN 


HEATER. 










(Temperature of steam, 212 degrees F.) 










Initial Temperature of Feed Water 


, Degrees F. 








40 


50 


60 


70 


80 


90 


100 


110 


120 


130 




2 


60.1 


69.9 


79.7 


89.5 


94.4 


109.2 


119.0 


128.8 


138.7 


148.5 


p 


3 


69.9 


79.6 


89.3 


90.1 


108.8 


118.6 


128.3 


138.0 


147.8 


157.5 


a 


4 


79.5 


89.1 


98.8 


108.5 


118.1 


127.8 


137.4 


147.1 


156.7 


166.4 


I s 


5 


89.0 


98.5 


108.1 


117.7 


127.2 


136.8 


146.4 


155.9 


165.5 


175.1 


;i 


6 


98.3 


107.7 


117.2 


126.7 


136.2 


145.7 


155.2 


164.7 


174.2 


183.6 




7 


107.4 


116.8 


126.2 


135.6 


145.0 


154.4 


163.8 


173.2 


182.5 


192.1 


o 2 


8 


116.4 


125.7 


135.0 


144.4 


153.7 


163.0 


172.4 


181.8 


191.0 


200.3 


O >j 


9 


125.2 


134.5 


143.7 


153.0 


162.2 


171.5 


180.7 


190.0 


199.2 


208.5 


t* 


10 


133.3 


143.1 


152.3 


161.4 


170.6 


179.8 


189.0 


198.1 


207.3 


212.0 


1 


11 


142.5 


151.6 


160.7 


169.7 


178.9 


188.2 


197.0 


206.2 


212.0* 


212.0* 


1 

di 


12 


150.9 


159.9 


168.9 


177.9 


187.0 


196.0 


205.0 


212.0* 


212.0* 


212.0* 



* All of the steam not condensed. 

Example : A power plant has 1200 I.H.P. of engines using 20 pounds 
of steam per I.H.P. hour. Auxiliaries use equivalent of 10 per cent of 
main engine steam. Pressure in heater pounds gauge, temperature 
of hot- well supply 110 degrees F. Required temperature of feed water 
leaving heater. 






FEED-WATER PURIFIERS AND HEATERS 491 

Here X = 1146 (from steam tables), t = 110, S = 0.10. 

Substituting these values in (105), 

0.9 X 0.10 (1146 - 1 + 32) - t - 110. 

t = 198 degrees F. 

253. Pan Surface Required in Open Feed- Water Heaters. — Pan or 

tray surface required varies according to the quality of the water with 
regard to both scale-making material and mud, and may be approxi- 
mated by the formula 

Pan surface, sq. ft. ^ Lb- of water heated per hr.X horsepower : (1Q5a) 





Vertical Type. 


Horizontal 
Type. 


For very muddy water, c 


118 
166 
500 


110 


Slightly muddy water, c 


155 


For clean water, c 


400 







254:. Size of Shell, Open Heaters. — General proportions of open 
heaters vary considerably on account of the different arrangements of 
pans or trays, filter and oil-extracting devices. A fair idea of the size 
of shell required may be obtained by the formulas 



. - , „ Horse power 

Area of shell = — — ; Sr~. — 7— ■ 

a X length in feet 



(106) 



Length of shell - Horsepower . 

a X area in square feet. 
a = 2.15 for very muddy water. 
a = 6 for slightly muddy water. 
a = 8 for clean water. 

The horse power in this case is obtained by dividing the weight of water 
heated per hour by the steam consumption of the engine per horse 
power per hour. 

Pans containing 2.5 square feet and less are usually made round, and 
larger sizes rectangular in plan. When circumstances will permit it 
is better to have not more than six pans in any one tier, since it is 
advisable to proportion the pans so as to obtain as low a velocity over 
each as practicable. 

Distance between trays or pans is seldom less than one-tenth the 
width for rectangular and one-fourth the diameter for round pans. 
Volume of storage and settling chamber in horizontal heaters varies 



492 



STEAM POWER PLANT ENGINEERING 



from 0.25 for good quality of water to 0.4 of the volume of the shell 
for muddy water, 0.33 being about the average. In the vertical type 
the settling chamber represents respectively 0.4 and 0.6 the volume of 
the shell with clear and muddy water. Filters occupy from 10 to 15 
per cent of the volume of the shell in the horizontal type and from 15 
to 20 per cent in the vertical type, the smaller percentage corresponding 
to clear water and the larger to muddy water or water containing a con- 
siderable quantity of impurities. 

Open Heaters: Cassier's Mag., Aug., 1903, p. 33; Engr. U.S., Jan. 1, 1906, pp. 17, 
78; St. Ry. Jour., Feb. 4, 1905, p. 227; Am. Elecn., Sept., 1905, p. 481. 



SURFACE BLOW 




EXHAUST FROM 

HEATER 



255. Classification of Closed 
Heaters. — Closed heaters may 
be grouped into two classes: 

1. Water tube, Fig. 242, and 

2. Steam tube, Fig. 246. 
Closed heaters, both water 

tube and steam tube may oper- 
ate with: 

1. Parallel currents, where the 
water and steam flow in the 
same direction, Fig. 242, or with 

2. Counter currents, where the 
water and steam flow in opposite 
directions, Fig. 244. 




Fig. 242. 



Goubert Single-Flow Closed 
Heater. 



Fig. 243. 



Details of Expansion Joint, 
Goubert Heater. 



Water-tube heaters may be still further classified as 
1. Single- flow, in which the water flows through the heaters in one 
direction only, Fig. 242. 



FEED-WATER PURIFIERS AND HEATERS 



493 



2. Multi-flow, in which the water flows back and forth a number of 
times, as in Fig. 244. 

3. Coil heater, in which the water flows through one or more coils, 
as in Fig. 245. 

256. Water-Tube, Closed Heaters. — Fig. 242 shows a section 
through a feed-water heater of the single-flow straight-tube type. The 



ot/rtEr 




Fig. 244. Wainwright Multi-Flow 
Closed Heater. 



Fig. 245. Typical Coil Heater. 



tubes are of plain brass and the shell of cast iron. The tubes are 
expanded into the tube sheets by a roller expander. To provide for 
expansion the upper tube sheet and water chamber are secured to the 
main shell by means of a special expansion joint the details of which 
are shown in Fig. 243. R is a ring or gasket of soft annealed copper 



494 STEAM POWER PLANT ENGINEERING 

and G, G two gaskets of special packing with brass wire cloth insertion. 
These gaskets form a flexible expansion joint between C and tube 
sheet D, so that the whole upper chamber, which is carried solely by 
the tubes, is free to move up and down as the tubes expand or contract 
under varying temperatures. 

Fig. 244 shows a section through a Wainwright heater, illustrating 
the multi-flow water-tube type. The body of the heater is of cast 
iron, the tubes of corrugated copper. The water passes through 
the tubes and the steam surrounds them. The feed water and 
exhaust steam do not mingle, and hence the oil in the exhaust 
does not contaminate the water. The water chambers are divided 
into several compartments, as shown in the illustration, and the par- 
titions are so arranged that the flow of feed water is directed back 
and forth through the various groups of tubes in succession. This 
arrangement gives a higher velocity of flow than the non-return 
type of heater, and therefore increases the rate of heat absorp- 
tion. The mud and impurities settle at the bottom and are 
discharged through the mud blow-off. Such impurities as rise to 
the surface are removed by the surface blow-off. The tubes are cor- 
rugated to allow for expansion and at the same time to increase the 
transmission of heat. Referring to Fig. 244: Exhaust steam enters 
at A and leaves at E, and the portion which is condensed is drawn off 
at D. Feed water enters at / and is discharged at 0. P, P are mud 
blow-offs and S is an opening for a safety valve. Table 66 gives results 
of tests showing the relative efficiencies of plain and corrugated tubes 
for various velocities. 

Fig. 245 shows a partial section through a Harrisburg feed-water 
heater. This apparatus is a typical example of the coiled-tube heater. 
Three sets of concentric copper coils are brazed to gun-metal manifolds 
and supported by clamp stays as indicated in the illustration. Feed 
water enters the heater at the bottom manifold and passes through 
the coils to the feed outlet. The exhaust steam enters the heater at 
the bottom and surrounds the coils in its passage to the outlet at the 
top. The coils are designed to withstand a pressure of 600 pounds 
per square inch. 

257. Steam-Tube, Closed Heaters. — Fig. 246 shows a section 
through an Otis heater, illustrating the steam-tube type. Here the 
exhaust steam passes through the tubes which are surrounded by the feed 
water. The exhaust steam enters at A, and passes down one section 
of tubes into the enlarged space of the water and oil separator 0, in 
which the condensation and oil are deposited. From this chamber the 
steam passes up through the other section of tubes to outlet C, thus 



FEED-WATER PURIFIERS AND HEATERS 



495 



passing twice through the entire length of the heater. The water enters 
at E and is discharged at G. R is the blow-off opening. The tubes 
are of seamless brass and are curved to allow for expansion. Condensed 
steam is withdrawn at P. 

Fig. 247 shows a partial section through a Baragwanath steam 
jacketed steam-tube heater. Exhaust steam enters at A, passes up 







#M Aft 



Fig. 246. Otis Steam-Tube Feed-Water 
Heater. 



Fig. 247. Baragwanath Steam-Jacketed 
Feed- Water Heater. 



through the tubes, returns down annular space E between the inner 
shell and jacket, and passes out at B. Feed water enters at C and 
leaves at D. E is the scum blow-off, G the heater drain, and H the 
jacket drain. 



496 STEAM POWER PLANT ENGINEERING 

258. Heating Surface, Closed Heaters. — It is generally assumed 
that the transfer of heat between two bodies is directly proportional 
to the difference in temperature between them. 

Let T = temperature of the water entering the heater. 

T 2 = temperature of the water leaving the heater. 

T s = temperature of the exhaust steam. 

A = square feet of transmitting surface. 

T = temperature of a unit of water t seconds after entering 
the heater. 

h = B.T.U. absorbed per square foot per second per degree 
difference in temperature between the steam tempera- 
ture T s and the water temperature T. 

t = time in seconds. 

w = number of pounds of feed water per second. 

Then — = square feet of surface brought in contact with one pound of 
w 

water per second, 

and dT, the rate at which the temperature of the water is increasing 

at this instant, will be 

dT= — (T 8 -T)dt. (108) 

w 



Integrating, 



pjL = Mf 4 (111) 

J T Ts-T w J 

loge TS ~ T ° = — • (112) 

Let W = number of pounds of feed water heated per hour. 

U = B.T.U. transmitted to the feed water per square foot of sur- 
face per hour per degree difference in temperature. 
Then (112) may be written 



from which 



x ^¥^rw' (113) 



A =v^T^k- (114) 






FEED-WATER PURIFIERS AND HEATERS 497 

Knowing the weight of water to be heated, the temperature of the 
s^eam, the desired temperature of the feed water, and the coefficient of 
heat transmission, U, this equation enables one to determine the area 
of heating surface required for the given conditions. Since the extent 
of heating surface increases rapidly as T 2 approaches T s , and becomes 
infinity for T 2 = T s , it is desirable to limit T 2 to some practical figure. 
An average maximum for T 2 = T s — 4. 

Table 64 has been calculated from this formula and gives the square 
feet of heating surface necessary to heat 1000 pounds of water per hour 
for different ranges in temperature. 

Mean Temperature Difference. 

If we let d = average temperature difference between the steam 
and feed water, then 

AUd = heat given out by the steam per hour. 
W (T 2 — T ) = heat absorbed by the feed water per hour. 
AUd = W(T 2 -T ). (115) 

(116) 
From (113), ^- = log, ^~ ^ . (117) 

Therefore d = T 2~ T ° . (118) 



Table 65 has been calculated from formula (118) and gives the 
mean temperature difference for various conditions of operation. 

The arithmetic mean temperature difference d t may be taken with 
safety for the average heater problem and has the advantage of sim- 
plicity. 

d t = T s ~ T «\ T * - (119) 

Closed heaters are sometimes rated on the basis of \ square foot of 
heating surface per horse power, i.e., a heater with 500 square feet of 
heating surface would be rated at 1500 horse power. 

259. Heat Transmission in Closed Heaters. — Table 66 gives the 
results of a series of tests on the absorption of heat by water passing 
through brass and copper tubes surrounded by steam. The curves in 
Fig. 248 were plotted from the data given in this table. An inspection 
of the table and the curves will show that the absorption of heat per 
square foot of surface per degree difference in temperature varies with 



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AU 




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= iog. r ;_ 


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T 2 


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r 2 -7 







iog. r ;_ 


■T. 



498 



STEAM POWER PLANT ENGINEERING 



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FEED-WATER PURIFIERS AND HEATERS 



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500 



STEAM POWER PLANT ENGINEERING 



the velocity of the water and the material and character of the tubes. 
Increasing the velocity of the water passing through the heater in- 
creases the rate of heat transmission and thereby renders the heating 

surface more effective. In order to 
employ moderately high velocities 
and at the same time allow suffi- 
cient time in which to raise the 
temperature to a maximum, the 
tubes should be as long as prac- 
ticable and of small diameters. 
Other things being equal, a heater 
containing a large number of tubes 
of small diameter is more economical 
than one containing a small number 
of large tubes. It is important to 
proportion the heater according to 
the amount of water to be heated 
and the maximum temperature to 
which the water must be raised. 
In designing a heater, then, the 
maximum amount of heat to be 
transmitted per degree difference in 
temperature per hour per square foot 
should be assumed, and the velocity 
of the water made such that it is 
capable of absorbing this amount. 
A good average figure for multi- 
"*5 50 75 ioo 125 150 175 fl ow heaters is U= 250 B.T.U. for 

plain brass or copper tubes and 
[/= 300 B.T.U. for corrugated tubes with a water velocity of 50 feet 
per minute; for single-flow heaters, U= 175 (for plain brass) with a 
water velocity of 12.5 feet per minute and for coil heaters U= 300 
(copper) with a water velocity of 150 feet per minute. These figures 
are for water-tube heaters only. For steam-tube heaters (iron tubes) 
a good average figure is U= 120. 

Experiments show that heaters and condensers operating with 
counter-currents are more efficient and are capable of obtaining a 
higher final temperature than those operating with parallel currents. 

Example: Determine the size of vacuum and atmospheric heaters 
for a condensing plant of 1200 I.H.P. Engines use 20 pounds of 
steam per I.H.P. hour; auxiliaries use the equivalent of 10 per cent 
of the main engine steam; vacuum 25 inches referred to 30-inch 




3 INDICATE PLAIN TUBES 

CORRUGATED TUBES 



FEED-WATER PURIFIERS AND HEATERS 501 

barometer; feed water, T = 50 degrees; temperature of hot well, 
T 2 = 110 degrees; coefficient of heat transmission, U= 300 B.T.U. 

Vacuum or Primary Heater. 
Feed water for main engines, 

20 X 1200 = 24,000 pounds per hour. 
Feed water used by auxiliaries, 

10 per cent of 24,000 = 2400 pounds per hour. 

Total feed, 

W= 24,000 + 2400 = 26,400 pounds per hour. 

From formula (114), 

W T S -T Q 
A = — - log e - 



U -=• T s -T 2 
26,400 lo 134 - 50 



300 ° 134-110 
= 110 square feet. 

On the basis of £ square foot of surface per horse power the rating of 
this heater will be 

110 X 3 = 330 horse power. 

Atmospheric or Secondary Heater. 

The temperature of the feed water leaving the atmospheric heater, 
formula (105), will be 

, _ t + 0-9 S (A + 32) 
% ~ 1+0.9S 

where S = .10, t = 110 degrees, X = 1146 B.T.U. 

whence t = 110 + 0.9x0-10(1146 + 32) 

1 + 0.9 X 0.10 
= 198 degrees. 

The required surface is 



U °°T S -T 2 

where T 8 = 212, T = 110, T 2 = 198, 

whence A = &™ log e 212 ~ 110 
300 6 212 - 198 

= 175 square feet. 
The horse-power rating will be 
175 X 3 = 525 horse power. 



502 



STEAM POWER PLANT ENGINEERING 



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504 STEAM POWER PLANT ENGINEERING 

260. Open vs. Closed Heaters. — Open and closed heaters have 
their respective advantages and a careful study of the various influ- 
encing conditions is necessary for an intelligent choice. The follow- 
ing parallel comparison brings out a few of the distinguishing features: 

Open Heater. Closed Heater. 

Efficiency. 

With sufficient exhaust steam for heat- The maximum temperature of the feed 
ing, the feed water may reach the water will always be 2 degrees or more 
same temperature as the steam. lower than the temperature of the 

Scale and oil do not affect the heat steam. 

transmission. Scale and oil deposit on the tubes and 

the heat transmission is lowered. 

Pressures. 

It is not ordinarily subjected to much The water pressure is slightly greater 
more than atmospheric pressure. than that in the boiler when placed 

on the pressure side of the pump as 
is customary. 

Safety. 

Sticking of the back pressure valve may It will safely withstand any pressure 
cause it to " blow up " if provision is likely to occur, 
not made for such an emergency. 

Purification. 

Since the exhaust steam and feed water Oil does not come in contact with the 

mingle, provision must be made for feed water. 

removing the oil from the steam. Scale is removed with difficulty. 

Scale and other impurities precipitated 

in the heater are readily removed. 

Location. 

Must always be placed above the pump May be placed anywhere on the pres- 
suction and on the suction side. sure side of the pump. 

Pumps. 

With supply under suction two pumps One cold-water pump is necessary, 
are necessary and one must handle 
hot water. 

Adaptability. 

Particularly adaptable for heating All vacuum or primary heaters are 
systems where it is desired to pipe necessarily of this type, 
the " returns " direct to heater. 



FEED-WATER PURIFIERS AND HEATERS 



505 



261. " Through " Heaters. — Fig. 249 shows a typical installation of 
a through heater in a non-condensing plant. 

It is evident that all the steam must pass through the heater. Now, 
one pound of exhaust steam in condensing gives up approximately 




Fig. 249. Open Heater Connected asa" Through " Heater. Non-Condensing Plant. 



1000 B.T.U. Hence, if the initial temperature of the feed water 
is 50 degrees and the final temperature 210, the engine furnishes 

— - — — = 6.26, say six times the quantity necessary for heating 
Z±k) — oU 

the feed water to a maximum. Therefore the area of the pipe 
supplying the heater with steam need be but 
one-sixth that of the main exhaust. With 
T D ^ ^ the heater connected as in Fig. 249 the connec- 
tions must necessarily be the same size as 
the exhaust pipe. 

With this arrangement the heater cannot be 
" cut out " while the engine is in operation and 
hence it is not adapted for plants working 
continuously. For the purpose of cutting out a 
heater while the plant is in operation a through 
heater may be by-passed as in Fig. 250. Ad- 
vantage may be taken here of the permissible 
reduction in the size of pipes and fittings, i.e., 
valves, etc., at C and D need be but one-half 
the size of those at A. This reduction in size may prove to be a con- 
siderable item in large installations. 




Fig. 250. 



506 



STEAM POWER PLANT ENGINEERING 



262. Induced Heaters. — Fig. 251 shows a typical installation of an 
induced heater in a non-condensing plant and Fig. 252 an induced 
primary heater in a condensing plant. 




Fig. 251. Open Heater Connected as an "Induced" Heater. Non-Condensing Plant. 



TO ATMOS 




ATMOSPHERIC 
RELIEF VALVE 



Fig. 252. Closed Heater Connected as an " Induced " Heater. Condensing Plant. 



In the arrangement in Fig. 251 the number of fittings is reduced 
to a minimum and the heater may be readily cut out. Since induced 
heaters are apt to become air-bound, a vapor pipe or vent is inserted 



FEED-WATER PURIFIERS AND HEATERS 507 

in the top of the heater as shown. This pipe varies from J to 1J 
inches in diameter,, depending upon the size of heater. 

Closed Heaters: Am. Elecn., May, 1900, p. 236, July, 1900, p. 354, Oct., 1905, 
p. 530; Cassier's Mag., Aug., 1903, p. 330; Eng. U.S., Jan. 1, 1906, p. 13; Power, 
April, 1902, p. 11. 

263. Live-Steam Heaters and Purifiers. — The function of a live- 
steam heater and purifier is primarily that of purification and hence it 
is not ordinarily installed unless the feed water contains scale-forming 
elements such as sulphates of lime and magnesia. These, as pre- 
viously stated, are not entirely precipitated until a temperature of 
approximately 300 degrees F. is reached; hence no amount of heating 
with exhaust steam at atmospheric pressure will thoroughly purify 
feed water containing these elements. 

Fig. 253 shows a section through a Hoppes live-steam purifier. 
Since the purifier is subjected to full boiler pressure, the shell and 



STEAM 




Fig. 253. Hoppes Live Steam Purifier. 

heads are constructed of steel. Within the shell are a number of 
trough-shaped pans or trays placed one above another and supported 
on steel angle ways. Steam from the boiler enters the chamber at A 
and comes in contact with feed water and condenses. The water on 
entering the heater at B is fed into the top pan and, overflowing the 
edges, follows the under side of the pan to the center and drops into the 
pan below. It flows over each successive pan in the same manner 
until it reaches the chamber at the bottom, whence it gravitates to the 
boiler through pipe C. As the steam inclosed in the shell comes in 
contact with the thin film of water, the solids held in solution are 
separated and adhere to the bottom of the pans in the same manner 



508 



STEAM POWER PLANT ENGINEERING 



that stalactites form on the roofs of natural caves. Authentic tests 
show that live steam heaters do not increase the boiler efficiency and 
that they merely act as purifiers. {Power, March 31, 1908, p. 498.) 
The purifier should be set in such a position as will bring the bottom 
of the shell two feet or more above the water level of the boilers, as in 
Fig. 254. N is the feed pipe from pump to purifier and should be pro- 




Fig. 254. Typical Installation of a " Live Steam" Purifier. 

vided with a check valve. D is the gravity pipe through which the 
purified water flows to the boiler. This pipe should be carried below 
the water level of the boilers and all branch pipes should be taken off 
below the water line. Pipe L leads from top of pipe S to pump 
or other steam-using device. This is necessary in order that air 
and other non-condensable gases liberated from the water may be 
removed from the purifier, which would otherwise become air-bound. 
In the illustration the feed pump takes its supply from an exhaust 
steam heater C. The purifier is provided with a suitable by-pass so 
that the water may be fed directly to the boiler when necessary. 

Live Steam Heated Feed Water: Elec. Engr., Lond., June 29, 1906; Am. Elecn., 
May, 1900, p. 214; Elec. Rev., Lond., May 20, 1898, p. 667; Eng. Rec, Aug. 30, 1898, 
p. 467; Power, March 31, 1908, p. 498. 

264. Economizers. — Fig. 256 gives a general view of a Green econo- 
mizer, illustrating a typical flue gas heater. It consists of a series of 
cast-iron tubes 9 to 12 feet in length and 4|-inches in diameter, which 



FEED-WATER PURIFIERS AND HEATERS 



509 




Fig. 255. Typical Installation of Primary and Secondary Heaters. 



510 



STEAM POWER PLANT ENGINEERING 







FEED-WATER PURIFIERS AND HEATERS 



511 



are arranged vertically in sections of various widths across the main flue 
between boiler and chimney. When in position the sections are con- 
nected by top and bottom headers, and the headers are connected 
to branch pipes running lengthwise, one at the top and the other at 
the bottom. Both of the branch pipes are outside the brickwork which 
incloses the apparatus. The waste gases are led to the economizer by 
the ordinary flue from the boiler to the chimney, but a by-pass must be 
provided for use when the economizer is out of service for cleaning or for 
repairs. The feed water is forced into the economizer through the lower 
branch pipe nearest the point of exit of gases, and emerges through the 
upper branch pipe nearest the point where the gases enter. Each 
tube is encircled with a set of triple overlapping scrapers which travel 
continuously up and down the tubes at a slow rate of speed, the object 
being to keep the external surfaces free from soot. The mechanism 
for working the scrapers is placed on top of the economizer, outside 
the chamber, and the motive power is supplied either by a belt from 
some convenient shaft or small independent engine or motor. The 
power for operating the gearing varies from 1 to J horse power per 
1000 square feet of economizer surface, depending upon the number 
and length of tubes. The apparatus is fitted with blow-off and safety 
valves, and a space is provided at the bottom of the chamber for the 
collection of soot. 

Fig. 257 shows a typical installation of a fuel economizer. 




Fig. 257. A Typical Economizer Installation. 



(Eng. Rec.) 



265. Value of Economizers. — The general conclusion drawn from 
current practice is that an economizer installation results in — 
A small annual saving in cost of operating the plant. 



512 STEAM POWER PLANT ENGINEERING 

Decreased wear and tear on the boilers due to the higher feed-water 
temperature. 

A large storage of hot water for sudden increase in load. 

Purification of the feed water due to the high temperature in 
the economizer. The scale-forming elements do not bake hard on the 
economizer tubes as they would in the boiler where the heat from the 
fire is more intense, but make a muddy deposit readily removed by 
blowing off. 

266. Factors Determining Installation of Economizers. — The factors 
to be considered before installing an economizer are: 

The nature of the auxiliary machinery, direct connected or belted. 

Method of heating the feed water; whether vacuum and atmos- 
pheric heaters are used and whether all or part of the auxiliary steam 
is used for heating. 

Initial temperature of the feed water; whether the feed is taken 
from the hot well or from a cold supply. 

Rise in temperature due to economizer. 

Cost of economizer. An approximate price is $15 per tube erected, 
on a basis of 15 square feet per tube. The heating surface is rated at 
3 to 5 square feet per boiler horse power. 

Cost of additional building space. 

Reduction in boiler-heating surface made possible by the econo- 
mizer. 

Extra cost of stack or forced-draft apparatus necessary to com- 
pensate for loss of draft due to economizer. The economizer lessens 
the draft by increasing the resistance between boilers and chimney 
and by reducing the chimney temperature. Where the installation of 
an economizer decreases the normal temperature of the chimney from 
say 550 degrees to 350 degrees F., the reduction in draft is approxi- 
mately 25 per cent. 

Total cost of economizer plant. This depends largely upon the 
design and varies from $4 to $7 per boiler horse power. 

Interest, depreciation, repairs, operation, taxes, and insurance. 

Table 67 gives the results of economizer tests. 

267. Temperature due to Use of Economizer. — The rise in temper- 
ature of feed water due to the use of an economizer may be approxi- 
mated from the following empirical formula advocated by the Green 
Economizer Company : 

x = y (^i - ^) H20) 

9l1 + 2 GC " 






FEED-WATER PURIFIERS AND HEATERS 



513 



in which 

x = rise in temperature of the feed water. 
T x = temperature of flue gas entering economizer. 
t x = temperature of feed water entering economizer. 
w = pounds of feed water per boiler horse power per hour. 
G = pounds of flue gas per pound of combustible. 
C = pounds of coal per boiler horse power per hour. 
y = square feet of economizer heating surface per boiler horse 
power. 




Fig. 258. 



Referring to Fig. 258, the ordinates represent temperatures, and 
abscissas the path of the flue gas and the water in the economizer. 
The flue gas enters the economizer at c with temperature T x and 
leaves at a with temperature T. The feed water enters at b with 
temperature t t and leaves at d with temperature t x + x. 

The algebraic mean temperature difference D between the flue gas 
and the feed water will be 

cd + ab 



D = 



T \ 


- (t t + x) + of— oa — bf 


T,- 


2 

- (t x + x) + T, — oa — U 



= T 1 -t l - 



2 

x + oa 



(121) 
(122) 
(123) 
(124) 



Now, wx = B.T.U. absorbed by the feed water per boiler horse 
power and 
GCS = B.T.U. given up to the feed water by the flue gas for 
each degree reduction in temperature (S = specific 
heat of the flue gas) ; therefore 
wx -r- GCS = total reduction in temperature of the flue gas; that is, 



wx 
GCS 



oa. 



(125) 



514 STEAM POWER PLANT ENGINEERING 

Substituting (125) in (124), we get 

wx 
fl-T.-f,- * ° CS (126) 

T t x GCS + w 
1 ' 2 0CS ' (127) 

in which 

D = mean temperature difference between flue gas and feed 
water, degrees F. 
Let U = B.T.U. absorbed per hour per square foot per degree 
difference in temperature and 
y = square feet of economizer surface per boiler horse power. 
Then UDy = heat absorbed per boiler horse-power hour. 
But wx = heat absorbed per boiler horse-power hour. 

Therefore wx = UDy. (128) 

Combining (128) and (127), 



wx = 



ti„(t / x_GCS_±jwx\ 

Uy^T^t- 2QCS ), (129) 

from which x = y(!Fl 7^W ( 13 °) 

w w + GCS 

U + 2 GCS 

y varies from 3.5 square feet to 5 square feet per boiler horse power, 
and U from 2.25 to 3.3, depending upon the conditions of operation.* 

If we let w = 30, S = 0.2, and U = 3.3, and substitute these values 
in equation (130), it assumes the form given by the Green Economizer 
Company, equation (120). 

A method of approximating the rise in temperature where the final 
temperature of the flue gas is known, is to assume T % degree rise in the 
feed water for each degree reduction in temperature in the flue gas. 
This is determined on the basis that approximately 20 pounds of flue 
gas are generated for each pound of combustible, and that 10 pounds 
of water are evaporated per pound of combustible; that is, 2 pounds 
of flue gas are generated for each pound of feed water delivered to the 
boiler. Assuming a specific heat of 0.25 for the flue gas, this gives 
2 X .25 or 0.5 degree rise in temperature in the feed water for each 
degree reduction in the flue gas temperature. 

Example: Determine the rise in temperature of the feed water in a 
power plant of 1200 I.H.P. Engines use 20 pounds of steam per 

* For D = 600 U = 3.25 For D = 400 U = 2.75 

500 3.00 300 2.25 






FEED-WATER PURIFIERS AND HEATERS 



515 



I.H.P. hour; auxiliaries use the equivalent of 12 per cent of main 
engine steam; vacuum 25 inches; feed- water supply 50 degrees; 3.7 
pounds of coal are burned per hour per boiler horse power; flue gas 
temperature 550 degrees F.; steam pressure 150 pounds gauge. 

The vacuum and atmospheric heater will raise the temperature of the 
feed water from 50 to 205 degrees. (See preceding problem.) 

On the assumption that 20 pounds of flue gas are generated per 
pound of combustible and that 3.5 square feet of economizer heating 
surface are installed per boiler horse power, the notations in the formula 
will become T t = 550, t x = 205, w = 30, G = 20, C = 3.7, y = 3.5, 
U = 3.3, S = 0.2. 

Substituting, 

3.5 (550-205) 

X 30 (5 X 30 + 20 X 3.7) 
3.3 (2X20X3.7) 

= 83 degrees rise in temperature. 
Therefore the temperature of the water entering the boiler will be 
205 + 83 = 288 degrees F. 

Economizers: Prac. Engr. U.S., March, 1910; Engr. U. S., March 15, 1900, p. 69; 
Power, July 27, 1909; Cassier's, March, 1900, p. 378; St. Ry. Jour., Oct. 31, 1903, 
1903, p. 822. 

Economizer Installations: Eng. Rec, March 15, 1902, p. 247, June 7, 1902, p. 532, 
Nov. 1, 1902, p. 410; Power, Feb., 1905. 

Tests of Economizers: Am. Elecn., Nov., 1902, p. 518; Elec. Rev., Lond., Aug. 
2, 1895, p. 148; Eng. Rec, March 9, 1901, p. 220, July 25, 1903, p. 102; Power, 
Aug., 1897, p. 8, Aug., 1904, p. 493; St. Ry. Rev., June, 1904, p. 436. 



TABLE 67. 

ECONOMIZER PERFORMANCES.* 









Temperatures. 






Fuel 
Saving. 


Plants 


Gases 


Gases 


Water 


Water 


Gain in 




Entering 


Leaving 


Entering 


Leaving 


Temperature 






Economizer. 


Economizer. 


Economizer. 


Economizer. 


of Water. 






Degrees. 


Degrees. 


Degrees. 


Degrees. 


Degrees. 


Per Cent. 


1 


610 


340 


110 


287 


177 


16.7 


2 


505 


212 


84 


276 


192 


17.1 


3 


550 


205 


185 


305 


120 


11.7 


4 


522 


320 


155 


300 


145 


13.8 


5 


505 


320 


190 


300 


110 


10.7 


6 


465 


250 


180 


295 


115 


11.2 


7 


490 


290 


165 


280 


115 


11.0 


8 


495 


190 


155 


320 


165 


15.5 


9 


595 


299 


130 


311 


181 


16.8 



* Transactions American Society of Mechanical Engineers. Vol. XV. 



516 . STEAM POWER PLANT ENGINEERING 

268. Choice of Feed-Water Heating System. — The heating of feed 
water and its delivery to the boiler in the most economical manner 
is a problem involving such a large number of combinations that a 
general analysis is impracticable. The following discussion of a spe- 
cific case will give some idea of the manner in which this problem may 
be attacked. 

Example: Determine the most economical manner of heating the 
feed water for a power plant of 1000 horse power operating under the 
following conditions: Schedule 10 hours per day and 310 days per 
year; load factor on the ten-hour basis 0.8; cost of coal $2.50 per 
ton of 2000 pounds; heat value of the coal 13,500 B.T.U. per pound; 
average boiler efficiency 65 per cent; engines use 20 pounds of steam 
per I.H.P. hour; steam pressure 150 pounds absolute; temperature of 
cold water 60 degrees; vacuum 26 inches referred to 30-inch barometer; 
interest 5 per cent; depreciation 8£ per cent; maintenance 1 per cent; 
insurance J per cent; taxes 1 per cent; total charges 16 per cent; 
charges for attendance and maintenance assumed to be the same in 
each case and credit for the chimney assumed to offset debit for econo- 
mizer space. Many of the influencing conditions are left out for the 
sake of simplicity. 

The most likely combinations are 

(1) Atmospheric, all auxiliaries steam driven, water taken from cold 

well. 

(2) Same as 1 except that water is taken from hot well. 

(3) Economizers, auxiliaries electrically driven, chimney draft, water 

from cold well. 

(4) Vacuum heater, economizer, and electrically driven auxiliaries, 

fan draft. 

(5) Vacuum heater, atmospheric heater, and steam auxiliaries. 

(6) Atmospheric heater, economizer, steam auxiliaries, fan draft. 

(7) Vacuum and atmospheric heaters, economizers, steam auxiliaries, 

and electrical fan. 

(8) Vacuum, atmospheric heater, economizer, and chimney draft, aux- 

iliaries operating condensing except feed pumps and stoker 
engines which exhaust into the atmospheric heater. 

The difference between the total heat furnished by the boiler and 
the heat returned in the feed water is the net heat put into the steam 
by the boiler. Evidently the system which shows the least net heat 
required to produce one horse power will be the most economical as 



FEED-WATER PURIFIERS AND HEATERS 517 

far as coal consumption is concerned, although not necessarily the 
cheapest when both operating and fixed charges are considered. 

Prices vary so much that it is practically impossible to give costs of 
installations which will bear criticism and the prices taken in this problem 
are approximate only. 

Case I. 

Atmospheric heater, auxiliaries steam driven, feed from cold well. 

This arrangement and that of Case II are the most common in power 
plants of this size. 

The power consumption of the auxiliaries operating non-condensing 
varies from 8 to 12 per cent of the total power developed. Assume 
it to be 10 per cent. 

The temperature of the feed water leaving the heater may be 
determined by formula (105). 

_ * + .9S(X + 32) 
1 + .9S 

Substituting S = .10, X= 1146, t = 60, 

t = 60 + -9 X .10 (1146 + 32) 
1 + .9 X .10 
= 152. 

The net heat furnished by the boiler to produce one indicated horse- 
power hour in the engine is evidently the heat necessary to raise 
20 + 10 per cent of 20 = 22 pounds of water from 152 degrees F. to 
steam at 150 pounds pressure; i.e., the net heat furnished is 

22 X 1071.2 = 23,564 B.T.U. 

Now, 1 I.H.P. = 2545 B.T.U. 

Therefore the heat efficiency of this arrangement is 

2545 1AC , 

10.8 per cent. 



23,564 



Probable First Cost. 

Steam pumps $400.00 

Condenser with steam-driven air and circulating pumps 3000.00 

1000 horse-power open heater 480.00 

Piping 1200.00 

$5080.00 



518 STEAM POWER PLANT ENGINEERING 

Fud Consumption. 

Average horse-power hours per year = 1000 (rated horse power) X 0.8 (curve 
load factor) X 310 (days per year) X 10 (hours per day) = 2,480,000. 

Pounds of coal per I.H.P. hour = net heat furnished per I.H.P. hour -f- net heat 
absorbed by the boiler per pound of coal = 23,564 ~ (13,500 X 0.65) — 2.68. 

Tons per year= 2,480,000X2.68 _ ^ 
2000 

Fuel and Fixed Charges. 

Fuel, 3323 tons at $2.50 $8,308.00 

Fixed charges, 16 per cent of $5080 , 812.00 



$9,120.00 
Case II. 

Same as Case I, except that feed is taken from the hot well. This 
arrangement is possible only when the condensing water is suitable 
for feed purposes. 

Assume the temperature of the water from the hot well as it enters 
the heater to be 110 degrees. 

The temperature of the feed water leaving the heater will then be 
198 degrees (from formula (105)). 

Net heat furnished = 22 X 1025.2 = 22,554 B.T.U. 
Efficiency = r^T^r = 11.3 per cent. 

Pounds of coal per I.H.P. = — f^ M „ = 2.62. 
r 13,500 X -65 

Tons per year= 2,480,000 X 2.62 = ^ 
F J 2000 

Fuel and Fixed Charges. 

Fuel, 3248 tons at $2.50 $8,120.00 

Fixed charges (same as Case I) 812.00 



$8,932.00 
Case III. 

Economizers, auxiliaries electrically driven, chimney draft, water 
from the cold well. 

Practice gives an average of 3 per cent of the main engine output as 
the power required to operate the electrical auxiliaries in a plant of this 



FEED-WATER PURIFIERS AND HEATERS 519 

The temperature of the feed water leaving the economizer may be 
determined from formula (120). 

y(T x -t x ) 



x = 



Q1 5wJ_GC 
9 ' 1+ 2GC V 



Substituting, 

3.5(550-60) 1in , 

* = Q1 . 5X30 + 20x ^ 7 - 119 degreeS ' 
2 X 20 X 3.7 

Temperature of feed water entering heater = 119 + 60 = 179 degrees. 
Net heat furnished = (20 + 3percent of 20) X 1044.2 = 21,510 B.T.U. 

Efficiency = ^^- = 11.8 per cent. 
y 21,510 F 

Probable First Cost. 

Economizers $3,500.00 

Motor feed pump 600.00 

Condenser with electrically driven air and circulating pump 6,000.00 

Piping and wiring 1,000.00 



$11,100.00 

Fuel Consumption. 

Pounds of coal per I.H.P. hour = - = 2.45. 

13,500 X .65 

_ 2,480,000 X 2.45 _.__ 

Tons per year = — - = 3038. 

2000 

Fuel and Fixed Charges. 

Fuel, 3038 tons at $2.50 $7,595.00 

Fixed charges, 16 per cent on $11,100 1,776.00 



$9,371.00 
Case IV. 

Vacuum heater, economizer, electrically driven auxiliaries, fan draft. 

The vacuum heater may be relied upon to raise the temperature of 
the feed water to 110 degrees. 

The economizer will increase this 107 degrees (from formula (120)), 
giving the feed water a temperature of 217 degrees as it enters the 
heater. 

The electrical fan for the mechanical-draft system will require approx- 
imately 2 per cent of the main system engine power, making a total of 
3 + 2 = 5 per cent for all auxiliaries. 



520 



STEAM POWER PLANT ENGINEERING 



Net heat furnished = (20 + 5 per cent of 20) X 1006.2. 
= 21,130 B.T.U. 



Efficiency = 



2545 
21,130 



= 12.05 per cent. 



Probable First Cost. 

For the sake of simplicity it is assumed that the high first cost of the 
chimney plus its low depreciation and maintenance will offset the low 
first cost of the mechanical-draft system plus its higher maintenance 
and depreciation charges. 

Economizers $3,500.00 

Motor feed pump 600.00 

Motor-driven pumps and condenser 6,000.00 

Motor-driven fan 750.00 

Piping and wiring 1,200.00 

Vacuum heater 200.00 

$12,250.00 

Fuel Consumption. 

Pounds of coal per I.H.P. hour = ■ == 2.41. 

13,500 X .65 

Tons per year = ?^500X^41 =2988> 
2000 



Fuel and Fixed Charges. 

Fuel, 2988 tons at $2.50 $7,470.00 

Fixed charges, 16 per cent of $12,250 1,960.00 

$9,430.00 

In like manner Cases V, VI, VII and VIII have been treated and are 
tabulated in the summaries. 

SUMMARY (1). 



I.... 
II... 
III.. 
IV.. 
V... 
VI.. 
VII. 
VIII 



Temperature 
of Feed 
Water. 



Degrees F. 
152 
198 
179 
217 
208 
294 
290 
270 



Power 
Consumed by- 


Efficiency. 


First 
Cost. 


Fuel Cost 
per Year. 


Auxiliaries. 








Per Cent. 


Per Cent. 






10 


10.8 


$5,080 


$8,308 


10 


11.3 


5,080 


8,120 


3 


11.8 


11,100 


7,595 


5 


12.05 


12,250 


7,470 


10 


11.4 


5,280 


7,900 


14 


12 


9,000 


7,750 


10 


12.2 


9,300 


7,380 


8 


12.3 


8,250 


7,075 



Cost of 
Operation 
per Year. 



$9,120 
8,932 
9,371 
9,430 
8,744 
9,190 
9,570 
8,395 



FEED-WATER PURIFIERS AND HEATERS 
SUMMARY (2). 



521 



Case. 


Efficiency. 


First Cost. 


Fuel. 


Cost per Year. 


I 


8 


1 


8 


4 


II 


7 


1 


7 


2 


Ill 


6 


6 


4 


6 


IV 


3 
5 


7 
2 


3 

6 


7 


V 


3 


VI 


4 


4 


5 


5 


VII 


2 


5 


2 


8 


VIII 


1 


3 


1 


1 







Summary (2) gives the ranking; thus: Case I is eighth in point of 
efficiency; first in cheapness of installation; eighth in yearly cost of 
fuel; and fourth in yearly cost of operation. Case VIII is apparently 
the best arrangement for the given conditions. 



CHAPTER XIII. 

PUMPS. 

269. Classification. — Pumps used in connection with steam power 
plants may be conveniently classified under five groups according to 
the principles of action. 

1. Piston pumps, in which motion and pressure are imparted to 
the fluid by a reciprocating piston, plunger, or bucket. The action is 
positive and a certain definite amount of fluid is handled per stroke 
under predetermined conditions of pressure and velocity. 

2. Centrifugal pumps, in which the fluid is given initial velocity and 
pressure by a rotating impeller. The action is not positive, as the 
amount of fluid discharged is not necessarily proportional to the impeller 
displacement. 

3. Rotary pumps, in which motion and pressure are imparted to 
the fluid by a rotating impeller. The volume discharged is practically 
equal to the impeller displacement regardless of pressure. 

4. Jet pumps, in which velocity and pressure are imparted to the 
fluid by the momentum of a jet of similar or other fluid. The ordinary 
steam injector is the best known of this group. 

5. Direct-pressure pumps, in which the pressure of one fluid acts 
directly on the surface of another fluid, thereby imparting all or part 
of its energy to the latter. The pulsometer is an example of this type. 

These groups may be variously subdivided as follows: 



Piston. 



Direct-acting.. \%^ 



Forcing. 
Lifting. 



^-heel j^ptf 

Power dnven •• I Triplex 

r< , ., , f Volute Single stage 

Centrifugal . . . . ( Turbine Multi-stage 

Rotary j Power driven . . j L ^ n g g 

T , ( T . , j Positive . . . 

Jet | Injector j Automatic 

Direct nressure \ P uls O m eter Lifting. . . . 

Uirect pressure -j Air _ Hft Lifting 

Piston or plunger pumps are the most common in use. Boiler- 
feed pumps, city waterworks pumps, and force pumps are ordi- 
narily of this type. In the direct-acting type, Fig. 260, the water 

522 



Air. 

Vacuum. 
Forcing. 
Lifting. 



PUMPS 523 

plunger and steam piston are secured to a single piston rod and the 
steam pressure is transmitted directly to the water. There is no fly- 
wheel, connecting rod, or crank. The velocity of the delivery is pro- 
portional to the resistance offered by the water; when the resistance 
equals the forward effort of the steam pressure the pump stops. This 
class of pump is well adapted for boiler-feeding purposes, since it 
may be operated as slowly as suits the requirements of feeding by 
simply throttling the discharge. The steam consumption is very 
large in proportion to the work performed, since the steam is not used 
expansively. 

Fly-wheel pumps, Figs. 273, 308, are ordinarily classified as pumping 
engines. In this class steam may be used expansively, as sufficient 
energy is stored in a fly wheel to permit the drop in steam pressure 
during expansion. These pumps find wide application in city water- 
works, elevator plants, and the like, where high duty is required. They 
are little used as stationary boiler feeders, but are used to some 
extent in river boat practice and in plants operating continuously for 
long periods at comparatively steady loads. Practically all sizes of 
dry-air pumps and a number of large jet condenser pumps are of this 
type. 

Piston pumps, Fig. 279, driven by gearing or belting are ordinarily 
classified as power-driven pumps. The driving power may be steam 
engine, electric motor, or gas engine. The single-cylinder machine is 
often designated as a " simplex " power-driven pump, the two-cylinder 
as a " duplex," the three-cylinder as a " triplex," and so on. 

Centrifugal pumps, Fig. 292, are supplanting to a considerable extent 
the present type of piston pump for many uses. Though particularly 
adapted for low heads and large volumes they are used in many 
situations requiring extremely high heads. They are not as efficient as 
high-grade pumping engines, but the extremely low first cost fre- 
quently offsets this disadvantage, and they are much used in connection 
with dry docks, irrigating plants, sewage systems, and as circulating 
pumps in condensing plants. 

Rotary pumps, Fig. 305, are employed to a limited extent in the 
same field as the centrifugal pump. Being positive in action, they 
permit of a much lower rotative speed for the same delivery pressure. 

Jet pumps, Fig. 282, are seldom used as pumps in the ordinary 
sense of the word, on account of their extremely low efficiency, 
but are frequently employed for discharging water from sumps. 
Their greatest field of application lies in boiler feeding and in this 
respect their efficiency is comparable with that of the average piston 
pump. 



524 



STEAM POWER PLANT ENGINEERING 



Direct-pressure pumps operated by steam, such as the " pulsometer," 
Fig. 309, are used principally for pumping out sumps, surface drains, 
and the like, where the operation is intermittent. Direct-pressure 
pumps of the air-lift type, Fig. 310, are quite common and are used a 
great deal in situations where water is to be pumped from a number 
of scattered wells. 

270. Boiier-Feed Pumps, Direct-Acting Duplex. — Figs. 259 and 260 
illustrate a typical duplex boiler-feed pump, which consists virtually of 



AIR 
CHAMBER 



DISCHARGE 



STEAM SUPPLY 




Fig. 259. Typical Duplex Pump. 

two direct-acting pumps mounted side by side, the water ends and the 
steam ends working in parallel between inlet and exhaust pipe. The 
piston rod of one pump operates the steam valve of the other through 
the medium of bell cranks and rocker arms. The pistons move alter- 
nately, and one or the other is always in motion, the flow of water 
being practically continuous. 

In general construction the steam pistons and valves are similar 
to those of steam engines. The valves in duplex pumps, however, 
have no lap. In order to reduce the valve travel to a minimum, 
and^till have sufficient bearing surface between the steam ports and 
the main exhaust ports to prevent the leakage of steam from one to 
the other, separate exhaust ports are provided which enter the cylinder 
at nearly the same point as the steam ports. This arrangement offers 



PUMPS 



525 



a simple means of cushioning the piston by exhaust steam, thus pre- 
venting it from striking the cylinder heads at the ends of the stroke. 
The valves of the duplex pump having no lap would, if connected 
rigidly to the valve stem, open one port as soon as the other had been 
closed, at about mid-stroke of the piston, thus cutting down the stroke 



01SCHAR4C 



L 



per 




Fig. 260. Section Through a Typical Duplex Boiler-Feed Pump. 

to about one-fourth the usual length. To obviate this difficulty the 
valves are given considerable lost motion by allowing sufficient clear- 
ance between the lock nuts on the valve stem; the latter, therefore, 
imparts no motion to the valve until the piston operating it has nearly 
completed the stroke. The lost motion between valves and lock 
nuts renders it impossible 
to stop the pump in any 
position from which it can- valve stem 
not be started by simply 
admitting steam, and 
therefore the pump has 
no dead centers. When 
one piston moves to the 
end of the stroke it pulls 
or pushes the opposite 
valve to the end of its travel; then when the piston starts back to the 
other end of its stroke the valve remains stationary, owing to the lost 
motion, until the piston has completed about one-half the stroke. 




J PISTON ROD 



Fig. 261. 



526 



STEAM POWER PLANT ENGINEERING 



VALVC STEM 




PISTON ROD 



During this time the opposite piston has completed a full stroke and 
the valve operated by it will have opened the steam port wide, so 

that while one valve covers 

M — jrV, _ - both steam ports the other 

Qf|:| M ; | I I fi'ltD is at the end of its travel. 

In some makes of pumps 
the stem is rigidly attached 
to the valves, the lost mo- 
tion being adjusted outside 
the steam chest as shown 
in Figs. 261 and 262, which 
represent two common constructions of duplex valve gear. 

Fig. 263 shows the valve and piston in the position occupied at the 
commencement of the stroke. 
At one end of the valve the 
steam port P is open wide and 
at the opposite end the exhaust 
port E is open wide. When the 
piston nears the opposite end of 
the stroke and reaches the posi- 
tion shown in Fig. 264 the steam 
escape through the exhaust port 
E is cut off by the piston, and 
since the steam port is closed, the 
remaining steam is compressed 
between the piston and cylinder head, thus arresting the motion of 
the piston gradually without shock or jar. 

The construction of the water 
end of single-cylinder and du- 
plex pumps is practically the 
same; any slight differences 
which may be found are con- 
fined to minor details which in 
no way affect the general design 
or operation of the pump. 
The piston is double acting, 
the single-acting cylinder being 
confined to power pumps or to 
steam pumps intended for very 
high pressures. In the old-style pumps it was the custom to use one 
large valve with a lift sufficient to give the required passage, but in 
modern practice the required area is divided among several small 




Fig. 263. 




Fig. 264. 



PUMPS 



527 



valves, so that each one is easily and cheaply removed in case of 
accident or wear, and slip is lessened. * 

The valves are carried by two plates or decks, the suction valves 
being attached to the lower plate and 
the delivery valves to the upper one, 
as shown in Fig. 260. 

The valves in practically all boiler- 
feed pumps are of the flat disk type, 
Fig. 265, held firmly to the seat by 
conical springs and guided by a bolt 
through the center. 

All pumps are provided with an air 
chamber on the discharge side, which 
acts as a cushion for the water, pre- 
vents excessive pounding, and insures Fig. 265. A Typical Pump Disk-Valve. 

a uniform flow. Fig. 266 shows a section through the steam end of 
a compound duplex pump. 





<mmtiA 




Fig. 266. Section Through Steam Cylinders of a Typical Compound Duplex Pump. 



271. Feed Pumps with Steam-Actuated Valves. — Single-cylinder 
direct-acting pumps, Fig. 267, are ordinarily operated by steam-actu- 
ated valves. The steam enters the chest C and passes to the left through 
the annular opening A formed between the reduced neck of the valve 
and the bore of the steam chest. It is thus projected against the 
inside surface of the valve head H before escaping through the port P 
and passing to the cylinder. Both the pressure and impulse due to 
velocity acting on the valve head H tend to close or restrict the 

* The modern Riedler pump is an exception. See Engineer, U.S., Nov. 15, 1907, 
p. 1040. 



528 



STEAM POWER PLANT ENGINEERING 



admission port by forcing the valve to the left. On reaching the 
cylinder and forcing the piston X toward the right, the pressure of 
the steam upon the opposite side of the valve head H is pressing the 
valve to the right, a movement which would give the admission more 
port opening at A and deliver more steam to the cylinder. The valve 
then holds a position depending upon the relative intensity of the two 
pressures, which tend to move it in opposite directions, the admis- 
sion steam, tending to close the valve, and cylinder steam, tending 




Fig. 267. Marsh Boiler-Feed Pump. A Typical Steam-Actuated Valve Gear. 



to open the valve wider. The steam valve, therefore, is always in a 
balanced position. The steam piston is grooved at the center, form- 
ing a reservoir for live steam R which is supplied from the upper 
chamber of the steam chest by passage E to the cylinder cap S, 
and thence by tube M and the hollow piston rod V. The steam 
in this annular piston space reverses the steam valve by pressing 
alternately against the outer surfaces of the valve heads H through the 
connecting passages 0, near each end of the cylinder. The tappets T 
are for the purpose of moving the valve by hand in case it fails to 



PUMPS 



529 




/~\ 



y~x 



move automatically. Steam-actuated valves are not as positive in 
action as mechanically operated valves, and hence are little used in 
situations where positive action is essential, as in fire-pump service. 

272. Air and Vacuum Chambers. — Air chambers in piston pumps 
are for the purpose of causing a steady discharge of water and of 
reducing excessive pounding at high 
speeds by providing a cushion for the 
water. The water discharged under pres- 
sure compresses the air in the air cham- 
ber somewhat above the normal pressure 
of discharge during each stroke of the 
water piston, and when the piston stops 
momentarily at the end of the stroke 
the air expands to a certain extent and 
tends to produce a uniform rate of flow. 

The volume of the air chamber varies 
from 2 to 3J times the volume of the 
water piston displacement in single- 
cylinder pumps, and from 1 to 2\ times 
in the duplex type. High-speed pumps are provided with air cham- 
bers of from 5 to 6 times the piston displacement. The water level 
in the air chamber should be kept down to one-fourth the height of 
the chamber. In slow-running pumps sufficient air may be carried 
into the pump chamber along with the water, but with high speeds a 
large part of the air will be discharged, and air must be forced into the 
chamber by mechanical means. The larger the chamber the more 
uniform will be the discharge pressure. 



KTM 



Fig. 268. Forms of Vacuum 
Chambers. 




Fig. 269. Different Arrangements of Vacuum Chambers. 

Vacuum chambers are frequently provided for the purpose of main- 
taining a uniform flow of water in the suction pipe and assisting in the 
reduction of slip. Such chambers should be of slightly greater volume 



530 



STEAM POWER PLANT ENGINEERING 



than the suction pipe and of considerable length rather than diameter. 
Fig. 268 illustrates two designs commonly used. The one in Fig. 268 (B) 
should be placed in such a position as to receive the impact of the 
column of water in the suction pipe as illustrated in Figs. 269 (A), 
269 (B), and 269 (C). The chamber illustrated in Fig. 268 (A) should 
be placed in the suction pipe below but close to the pump. 

273. Water Pistons and Plungers. — In cold-water pumps the water 
pistons are usually packed with some kind of soft packing. Fig. 270 (A) 




(A) (B) 

Fig. 270. Types of Water Pistons. 



(c) 



shows the details of a piston with square hydraulic packing. The body E 
is fastened to the piston rod by nut C; packing is placed at D, and 
follower F is forced up by the nut B and locked by nut A. For large 




Fig. 271. Plunger with Metal Packing Pang. 

sizes the design is the same except that the follower is set up by a 
number of nuts near the edge. In hot-water pumps the pistons are 
often packed by means of metallic piston rings R, R, Fig. 270 (C), similar 
to those in steam pistons, or merely by water grooves G, G, Fig. 270 (B). 



PUMPS 



531 



The water end is often fitted with a plunger instead of a piston, as in 
Figs. 271 to 273. The piston is more compact, but the plungers do 
not require a bored cylinder, so that the first cost is not materially 
different. 

Fig. 271 shows a plunger with metal packing ring. When leakage 
becomes excessive it is necessary to renew the ring, which is readily 
removed. 

In Fig. 272 the plunger is packed with hydraulic packing as in the 
follower type of pump piston. The great difficulty with the above 




Fig. 272. Plunger with Hydraulic Packing. 



types of piston and plunger is in keeping the packing tight or in know- 
ing when it is leaking, and the trouble necessary to replace the packing. 
The outside packed plunger, Fig. 273, obviates these disadvantages to a 
great extent, since leakage is readily detected and repacking is performed 
without removing the cylinder heads. In dirty, dusty locations, how- 
ever, the piston pump or inside packed plunger is to be preferred, 
since the abrasive action of the dust renders outside packing difficult. 
Fig. 273 illustrates a high-duty elevator pump with outside packed 
plunger. 

274. Performance of Piston Pumps. — Direct-acting pumps as a 
class are wasteful of fuel and low in efficiency, due largely to the non- 
expansive use of steam. The average small duplex boiler-feed pump 
uses from 100 to 200 pounds of steam per I.H.P. hour, depending 
upon the speed, and the mechanical efficiency varies from 50 per cent 



532 



STEAM POWER PLANT ENGINEERING 




E3 

o 

+3 

I 

Ah 





* 



PUMPS 



533 



to 90 per cent. When new and in proper working condition the 
mechanical efficiency is seldom less than 85 per cent; but such pumps, 
as a rule, are given scant attention, and the average efficiency is not 
far from 65 per cent. The term "mechanical efficiency" in this 
connection refers to the ratio of the actual water horse power to 
the indicated horse power of the steam cylinder. The loss includes the 
slip of the piston and valves. A steam consumption of 150 pounds per 
I.H.P. hour with mechanical efficiency of 65 per cent is equivalent to 
a power consumption of about 5 per cent of the rated boiler capacity, 
although if the exhaust steam is used for feed-water heating the actual 
heat consumption may be but 1 to 1.5 per cent. Compound direct- 
acting pumps running non-condensing use from 50 to 100 pounds of 
steam per I.H.P. hour. Single-cylinder fly-wheel pumps of the slow- 
speed type, running non-condensing, use about 50 pounds of steam 
per I.H.P. hour. Multi-cylinder fly-wheel pumps of the high-duty 
type use about 25 pounds per I.H.P. hour when running non-condens- 
ing, and as low as 10 pounds when operating condensing. High-grade 
direct-connected motor-driven power pumps have a mechanical efficiency 
from line to water load, at normal rating, of about 80 per cent. The 
efficiency of geared pumps at normal rating varies with the character of 
the gearing and the degree of speed reduction, and may range anywhere 
from 40 to 70 per cent. 



400 



a 200 



I ioo 
a 

























Effect of Speed 
on the 
Economy of Small Direct- Acting - 
Steam Pumps 










A 16 x 
B 12 x 


10 x 12 

r^x 12 


Duplex 
Simplex 








\b 








- - 

















— * A — 


• 











50 75 100 125 150 

Number of Single Strokes JPer_Hinute 

Fig. 274. 



175 



20C 



The steam consumption of all direct-acting boiler pumps decreases 
with the increase in speed. This is illustrated by curve B, Fig. 274, 
plotted from the tests of a 12 x 7J x 12 direct-acting single-cylinder 
pump at Armour Institute of Technology, and curve A based on experi- 



534 



STEAM POWER PLANT ENGINEERING 



merits with a 16 x 12 duplex fire pump at Massachusetts Institute of 
Technology. 

Fig. 275 gives the details of the performance ofal2x7Jxl2 Marsh 
boiler-feed pump at the Armour Institute of Technology. 



vuuu 












Curves of Performance 














/ 














OI 

Marsh Steam Pump 

for 

Varying Speed 

Size of Pump— 12"x l^'x 12" 

Cap. 216 Gal. per Min. at 100 Strokes 








z 


/ 


























6000 














































1 
















































































5000 
c 




\ 
























/A 














\ 






















/£ 










r u 






v 




















/6 
















u 




i 


d 


































■a 


Pi ■ 
£•4000 

M 






Y& 






























u 




1 


Of 




\* 




























g 




o 
= 


0J 


■\ 






\ 










A& 














g 


15 


3000 


2 \ 

03 


\ 






<> 


























w 




H 


£ 


\ 
































a 


§1 




\ 






























& 


^ 






\ 








1/ 






















X 


X 




































^ 


y 






a 


2000 








































§ 
































t 




















^ 


fe 








^ 


















/ 












<% 


e y 


st* 




















1000 




f 




































/ 










































/, 










































// 










































k> 









































20 30 40 50 60 TO 
Single Strokes per Min. 

Fig. 275. 



5 S 



90 



100 



The determination of the power consumption of a boiler-feed pump 
is best illustrated by the following example. 

Example: A small direct-acting duplex pump uses 150 pounds of 
steam per I.H.P. hour. Gauge pressure 150 pounds per square inch; 
feed-water temperature 64 degrees F. Required the per cent of rated 
boiler capacity necessary to operate the pump. 

The head pumped against, 150 pounds per square inch, is equivalent 
to 150 X 2.3 = 345 feet of water. 






PUMPS 535 

The friction through the valves, fittings, and pipe, and the vertical 
distance between suction and feed-water inlet, are assumed to be equiva- 
lent to 20 per cent of the boiler pressure, giving a total head of 150 + 
30 = 180 pounds per square inch, or 414 feet of water. 

A boiler horse power, taking into consideration leakage losses and 
the steam used by the feed pump, will be equivalent to the evapora- 
tion of approximately 32 pounds of water per hour from a feed tem- 
perature of 64 degrees F. to steam at 150 pounds gauge. 

The actual work done in pumping 32 pounds of water against a head 
of 414 feet is 

414 X 32 = 13,248 foot-pounds. 



This corresponds to 

13,248 



= 0.0067 horse power. 



60 X 33,000 

The total heat of one pound of steam above 64 degrees F. is 1161.2 
B.T.U. The heat delivered to the pump per I.H.P. hour is 

1161.2 X 150 = 174,180 B.T.U. 

The amount used by the pump for each boiler horse power, disregard- 
ing efficiency, is 

174,180 X 0.0067 = 1167 B.T.U. per hour. 

The mechanical efficiency of the average feed pump ranges from 50 
to 85 per cent, depending upon its condition and the number of strokes 
per minute. Assuming it to be 65 per cent, the heat used by the pump 
per hour to deliver 32 pounds of water into the boiler is 

1167 4- 0.65 = 1795 B.T.U. 

A boiler horse power is equivalent to 33,320 B.T.U. per hour. There- 
fore the per cent of boiler output necessary to operate the pump is 

100X 3lS= 5 - 4perCent - 

If the exhaust steam is used for heating the feed water, the steam con- 
sumption will be 1.37 per cent of the boiler capacity, thus: The weight 
of steam consumed per boiler horse-power hour 

1795 



1161.2 



= 1.54 pounds. 



Allowing 10 per cent for condensation, the heat in the exhaust avail- 
able for heating the feed water is 

966 X 0.90 X 1.54 = 1340 B.T.U.* 

* Surface Condenser Plant. 



536 STEAM POWER PLANT ENGINEERING 

1795 — 1340 = 455 B.T.U., or the net heat required by the pump 
per hour to deliver 32 pounds of water to the boiler. 

The per cent of boiler output necessary to operate the pump is 

100 455 = 1.37. 
33,320 

Pump performances are generally given in terms of the foot-pounds 
of work done by the water piston per thousand pounds of dry steam or 
per million B.T.U. consumed by the engine, thus: 

1. Duty - Foot-pounds of work done x 1(m 

Weight of dry steam used 

2. Duty= , Foot-pounds of work jong x 1;00 0,000. 

Total number of heat units consumed 

(See A.S.M.E. code for conducting duty trials of pumping engines, 
Trans. A.S.M.E., 12-530, 563.) 

Example: A compound feed pump uses 100 pounds of steam per 
I.H.P. hour; indicated horse power, 48; capacity, 400 gallons per 
minute; temperature of water, 200 degrees F.; total head pumped 
against, 175 pounds per square inch; steam pressure, 100 pounds gauge; 
moisture in the steam, 3 per cent. Required the duty on the dry steam 
and on the heat-unit basis. 

175 pounds per square inch is equivalent to 175 X 2.4 = 420 feet of 
water at 200 degrees F. 

Weight of 400 gallons of water at 200 degrees F. = 400 X 8.03 = 
3212 pounds. 

Work done per minute = 3212 X 420 = 1,349,040 foot-pounds. 

Weight of dry steam supplied per minute = X 0.97 = 77.6 

pounds. 

B.T.U. supplied per minute = 10 ° * 48 (0.97 X 876.2 + 308.8-200 

+ 32) = 79,256. 
Duty per thousand pounds of dry steam 

= 1 > 349 » 040 x 1000 = 17,384,150 foot-pounds. 
77.6 

Duty per million B.T.U. 

= i; 349 ? 040 x 1000,000 = 16,893,863 foot-pounds. 
79,256 ' ' F 

Table 68 may be used in approximating the duty, thus : 

The mechanical efficiency of the pump in the preceding problem is 



PUMPS 



537 







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538 



STEAM POWER PLANT ENGINEERING 



At the intersection of vertical column " 85 " and horizontal column 
100 of Table 68, we find 16.82 millions. 

Tables 69 and 70 give the maximum theoretical height to which 
pumps may lift water by suction at different temperatures. In prac- 
tice these figures cannot be realized. It is customary to have the water 
gravitate to the pump for all temperatures over 120 degrees F. 

TABLE 69. 

MAXIMUM HEIGHT TO WHICH PUMPS CAN RAISE WATER BY SUCTION. 
(Temperature of Water 40 degrees F.; Barometer 29.92.) 



Vacuum in 

Suction Pipe, 

Inches of 


Theoretical 
Lift. 


Probable 

Actual 

Lift. 


Vacuum in 

Suction Pipe, 

Inches of 


Theoretical 
Lift. 


Probable 

Actual 

Lift. 


Mercury. 




Mercury. 






Feet. 


Feet. 




Feet. 


Feet. 


1 


1.1 


0.9 


16 


18.0 


14.4 


2 


2.2 


1.8 


17 


19.1 


15.3 


3 


3.3 


2.7 


18 


20.2 


16.1 


4 


4.5 


3.6 


19 


21.4 


17.1 


5 


5.6 


4.5 


20 


22.5 


18.0 


6 


6.7 


5.4 


21 


23.7 


18.9 


7 


7.9 


6.3 


22 


24.8 


19.8 


8 


9.0 


7.2 


23 


25.9 


20.7 


9 


10.1 


8.1 


24 


27.0 


21.6 


10 


11.3 


9.0 


25 


28.2 


22.7 


11 


12.4 


9.9 


26 


29.3 


23.9 


12 


13.5 


10.8 


* 27 


30.4 


24.3 


13 


14.6 


11.7 


28 


31.6 


25.2 


14 


15.8 


12.6 


29 


32.7 


26.1 


15 


16.9 


13.5 


f 29.68 


33.6 









* Vacua greater than 27 inches are practically unobtainable in pumping practice except in 
connection with condensers. 

t Maximum theoretical vacuum obtainable with water at 40 degrees F. and barometer of 
29.92 inches. 

TABLE 70. 

MAXIMUM THEORETICAL HEIGHT TO WHICH A PUMP CAN LIFT WATER BY 
SUCTION AT DIFFERENT TEMPERATURES. 



(Barometer 29.92.) 



Temperature of 


Maximum 


Temperature of Feed 


Maximum 


Feed Water. 


Theoretical Lift. 


Water. 


Theoretical Lift. 


Degrees F. 


Feet. 


Degrees F. 


Feet. 


40 


33.6 


130 


29.2 


50 


33.5 


140 


27.8 


60 


33.4 


150 


25.4 


70 


33.1 


160 


23.5 


80 


32.8 


170 


20.3 


90 


32.4 


180 


16.7 


100 


31.9 


190 


12.8 


110 


31.3 


200 


7.6 


120 


30.3 


210 


1.3 



PUMPS 539 

275. Size of Boiler-Feed Pump. 

Let D = diameter of water cylinder, inches. 

d = diameter of the steam cylinder, inches. 

L = length of stroke, inches. 

N = number of working strokes per minute. 

H = head in feet between suction and boiler water level. 

R = resistance in pounds per square inch between suction level 
and boiler water level due to valves, pipes, and fittings. 

p = boiler pressure, pounds per square inch. 

S = ratio of the water actually delivered to the piston displace- 
ment. 

W = weight of water delivered, pounds per hour. 

I = indicated horse power of the pump at maximum capacity. 

E = mechanical efficiency of the pump, taken as the ratio of the 
water horse power at the discharge opening to the indicated 
horse power of the pump, steam end. 



Then 



W = 2.Z..Mx60X 62.5 X S = 1.7 D 2 LNS. (132) 
4 144 12 



z> = °- 77 VlI^- < 133) 



^D ^ + fi + Q ll, (134) 

/ = w (p + R + °- 433 H ) 2 - 3 nw 

33,000 X 60 X E K 

In average practice the piston or plunger displacement is made about 
twice the capacity found by calculation from the amount of water 
required for the engine, to allow for leakage, steam consumption of the 
auxiliaries, blowing off, and pump slip. 

For pumps with strokes of 12 inches or over, the speed of the plunger 
or piston is usually limited to 100 feet per minute as a maximum to insure 
smooth running. "For shorter strokes a lower limit should be used. The 
maximum number of strokes ranges from 100 for strokes over 12 inches 
in length, to 200 for strokes under 5 inches. Boiler-feed pumps should 
be designed to give the desired capacity at about one-half the maximum 
number of strokes or less. 

Pump slip varies from 5 to 40 per cent, depending upon the condition 
of the piston and valves and the number of strokes. An average value 
for piston and plunger pumps in first-class condition is 8 per cent when 



540 STEAM POWER PLANT ENGINEERING 

operating at rated capacity, but it is wise to allow a much larger figure, 
say 20 per cent, for leakage caused by wear. 

The area of the steam cylinder is made from 2 to 2.5 times that of 
the water end to allow for the various friction losses and the drop in 
pressure between the pump throttle and the boiler. The total head 
pumped against includes the suction lift, the friction of valves and 
fittings, the distance between the suction inlet and the boiler level, and 
the boiler pressure. The excess head varies in practice from 15 to 40 
per cent of the boiler pressure; an average figure is 25 per cent. In 
allowing for the drop in steam pressure between boiler and pump a liberal 
figure is 25 per cent. 

The application of formulas (132) to (135), including the practical 
considerations stated above, is best illustrated by a specific example. 

Example: Determine the size of direct-acting single-cylinder feed 
pump necessary to supply water to 1000 horse power of boilers. Gauge 
pressure 100 pounds per square inch; feed-water temperature 150 
degrees F. 

One horse power is equivalent to the evaporation of 34.5 pounds of 
water from and at 212 degrees F.; but the pump is usually designed to 
supply about twice the capacity. 

Thus W = 62,400 (under the given conditions). 
S = 0.8 (by assumption). 
LN = 1200 (on the basis of 100 feet per minute). 

Substitute these values in (133) : 

D = 0.77 J 62,4Q( lo = 6 - 2 inches, — call it 6 inches, 
V 1200 X 0.8 

since the assumptions have been very liberal. 
Assume (.433 H + R) = 0.25 p and E = 0.65. 
Substitute these values in (134) : 



d = *\ll 



100 + 25 



0.65 X 100 
= 8.35, — call it 8.5 inches. 
Allowing 100 strokes per minute the length of the stroke must be 

L = 1200 -^ 100 = 12 inches. 
The dimensions of the pump are 8J x 6 x 12. 

The indicated horse power at maximum load may be obtained by 
substituting the proper values in (135), thus: 

j m 62,400 (100 + 25) 2.3 
33,000 X 60 X 0.65 
= 13.9 I.H.P. 



PUMPS 



541 



276. Steam-Pump Governors. — Fig. 276 shows a section through a 
Fisher pump governor, illustrating a device for maintaining a practically 
constant pressure in the discharge pipe irrespective of the quantity of 
water flowing. It embodies a pressure-reducing valve in the steam 

e supply pipe of the pump, actuated by the slight 

variations in water pressure. When the demand 
for water increases, the pressure in the discharge 
pipe tends to decrease, and this drop in pressure 
(transmitted to the pump governor by suitable 
piping) causes more steam to be admitted, which 
increases the speed of the pump. The governor 
is connected to the steam inlet of the pump at 
B and the steam enters at A. Double-balanced 
valve C regulates the supply of steam to the 
cylinder by the amount it is raised from the 
seat. The valve is held open by spring G, the com- 
pression of which may be regulated by hand wheel 
K. The water pressure from the discharge pipe 
acts on piston F and tends to overcome the resist- 
ance of the spring. The difference in pressure 
between the water and the spring determines the 
position of valve C. 

Piston rod H is pinned to sleeve I and valve 
stem L screwed into this sleeve by means of hand 
wheel K. Hence during ordinary operation the 
piston, piston rod sleeve, valve stem, and valve 
act as a single unit. By turning the hand wheel 
K, valve stem L will screw into sleeve / and the 
tension on the spring will be increased. Hand wheel J serves as a 
lock nut and prevents K from turning during normal operation. 

277. Feed- Water Regulators. — The water level in the boiler should 
be kept as nearly constant as possible, and this necessitates considerable 
attention on the part of the fireman, especially with fluctuating loads. 
There are a number of devices on the market which are designed to 
automatically maintain a constant level, and in many small plants 
where the duties of the fireman are numerous such devices in connec- 
tion with high and low water alarms are of considerable assistance. 
Their action, however, is not always positive on account of wear or 
sticking of parts, and engineers as a rule prefer to rely upon hand regula- 
tion. In large stations regulators are seldom used. 

Fig. 277 shows a section through a Kitts feed-water regulator, con- 
sisting of two parts, the chamber F and the regulating valve V. The 




Fig. 276. Fisher Pump 
Governor. 



542 



STEAM POWER PLANT ENGINEERING 



float chamber is connected to the boiler or water column at and E, 
and the regulating valve to the feed main at R and to the boiler feed 
pipe at W. When the water in the boiler falls below the mean level, the 
weight B overcomes the counterweight G and closes needle valve L by 
means of compound levers. At the same time an extension on valve L 
lifts spring A and opens exhaust valve D. This removes the steam 





Fig. 277. Kitts Feed-Water Regulator. 



Fig. 278. Rowe Feed-Water 
Regulator. 



pressure from the top of diaphragm C, in the regulating valve, through 
the agency of pipe K. The pressure from the pump raises the disk T 
and water flows into the boiler until the water rises to the mean level. 
When weight B becomes submerged its weight is overcome by counter- 
weight G, valve L is opened and exhaust valve D is closed. This 
admits steam pressure to the diaphragm C and forces disk T to its seat, 
cutting off the supply of water to the boiler. 

The Rowe feed-water regulator, Fig. 278, depends for its operation 
on a familiar float-controlled valve mechanism. The vessel A is con- 
nected to the boiler above and below the water line, and the float C, 
following the water level up and down, actuates a balanced valve in 
accordance with the boiler-feed requirements. When this apparatus 
is used to regulate the feed of a single boiler the opening G in the valve 
chamber is connected to the steam space of the boiler and the outlet H 



PUMPS 



543 



is carried to the steam inlet of the feed- water pump. When the water 
level is normal the float closes the valve L and thereby cuts off the 
supply of steam to the pump cylinders. Communication between 
chambers A and R is prevented by means of a diaphragm M. When 
the water level falls below normal the float pulls the valve down, open- 
ing the way for steam to pass from the inlet G to the outlet H and 
thence to the pump. When the regulator is used to control a battery 
of boilers the pump discharge delivers into the inlet G and the water 




Fig. 279. A Typical Geared Triplex Pump. 



passes through H to the boiler-feed main. Should the water level fall 
beyond a predetermined limit by reason of any accidental discontinuance 
of the water supply which the apparatus cannot correct, the float would 
open the valve F of the alarm whistle mounted on the top of the 
main vessel. 

278. Power Pumps. — Piston pumps, geared, belted, or direct con- 
nected to electric motors, gas engines, and water motors, are used 
chiefly where steam power is not available. Their general utility is 
evidenced by the rapidly increasing number installed in situations 
formerly occupied by the direct-acting steam pump. The efficiency of 



544 



STEAM POWER PLANT ENGINEERING 



this type of pump depends in a large measure upon the character of 
the driving motor and the efficiency of the transmitting mechanism. 
High-speed power pumps direct connected to electric motors give 































































































































































Knowles High Speed Electric Pump 

Direct connected to 

M P 6-100 H.P.-280-220 V.Form L 

Load and Efficiency 


































































































































































































































1 




1 


I 


























































Noz 


de 


Hprizoi 


ttal 












































































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/ 


































































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1 


















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4& 






















G 


all 


on 


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-2C 


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1 

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& 


































































v/ 




































































A 




































































J 


























10- 












































































-ic 




























































-1 


)0- 


3 


-200- 








































Pl 














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-9 






























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1 










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80 


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10 


u 


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y 


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100 






200 






300 






400 






500 






600 




























i 






1 






I 






1 






| 






1 



















Gauge Pressure at Valve (Lb.) 
Fig. 279a. 



efficiencies from line to water horse power as high as 83 per cent, while 
the low-speed geared type seldom exceed 70 per cent. The curves in 
Fig. 279a give the performance of a direct-connected triplex pump, 






PUMPS 



545 



and those in Fig. 280 the performance of a triplex pump geared to an 
electric motor. Both of these performances are exceptionally good 
and are considerably above the average. 

Power Pumps: Eng. Mag., Jan., 1905, p. 616; Col. Guard, Nov. 17, 1905; Elec. 
Rev., Jan. 10, 1902; Elec, N.Y., Oct. 12, 1904; Elec. World, Oct. 14, 1905, p. 667; 
Engr., Lond., Oct. 17, 1902, p. 377; Engineering, Sept. 1, 1905, p. 275; Engr. U.S., 
Jan. 1, 1904, p. 47. 



1UU 










Motor _ 
Pump 








md Gearin 




0> 
















Set 






Mechanical Efficienc 
Per Cent 

8 8 
















Head Cons 


ant, 350 Ft. 



















200 400 600 800 1000 

Discharge, Gallons Per Minute 



1200 



1400 



Head Constant. Speed Variable. 



100 



I 

P 



O 50 









Motor 










puroP_3 


id Gearing 
















Set 














y 






10 x 


2 Triplex P 
5 H.P. Mote 


imp 
r 










Mc 
C 

Capacit 


tor B.P.M. 

(earing 10 to 
y 1250 Gal. p 


150 

jr Min. 



25 50 75 100 125 

Total Head, Lb. Per Sq. In. Gauge 
Speed Constant. Head Variable. 



150 



.175 



Fig. 280. Performance of a 65-Horse-Power, Motor-Driven Triplex Pump. Geared Type. 

279. Injectors. — As a boiler feeder the injector is an efficient 
and convenient device, cheap and compact, with no moving parts, 
delivers hot water to the boiler without preheating, and has no exhaust 



546 



STEAM POWER PLANT ENGINEERING 



steam to be disposed of. Its adoption in locomotives is practically 
universal, but in stationary practice it is limited to small boilers or 
single boilers or as a reserve feeder in connection with pumps. The 
objections to an injector are its inability to handle hot water, the 
difficulty of maintaining a continuous flow under extreme variation of 
load, and the uncertainty of operation under certain conditions. Fig. 
281 illustrates the simplest form of single-tube injector. Boiler steam 



STEAM SUPPLY 




Fig. 281. Elementary Steam Injector. 

is admitted at A and, flowing through nozzle and combining tube to 
the atmosphere through G, partially exhausts the air from pipe B, 
thereby causing the water to rise until it comes in contact with the steam. 




Overflow 
Fig. 282. Hancock Double-Tube 
Injector. 




Fig. 283. 



Penberthy Automatic 
Injector. 



The steam emerging from nozzle C at high velocity condenses on meeting 
the water and imparts considerable momentum to it. The energy in 



PUMPS 547 

the rapidly moving mass is sufficient to carry it across opening 0, lift 
check H from its seat and force it into the boiler. The steam then 
ceases to escape at G. 

280. Positive Injectors. — Fig. 282 shows a section through a Han- 
cock injector, illustrating the principles of the double-tube positive 
type. Its operation is as follows: Overflow valves D and F are 
opened and steam is admitted, which at first passes freely through the 
overflow to the atmosphere and in so doing exhausts the air from the 
suction pipe. This causes the feed water to rise until it meets the jet 
of steam and the two are forced through the overflow. As soon as 
water appears at the overflow, valve D is closed, valve C partially 
opened, and valve F closed. This admits steam through the forcing 
jet W and, the overflow valves being closed, the water is fed into the 
boiler. In case the action is interrupted for any reason it is necessary 
to restart it by hand. 

The chief advantage of the double-tube positive type lies in its ability 
to lift water to a greater height and to handle hotter water than the 
single-tube. Its range in pressure is also greater, that is, it will start 
with a lower steam pressure and discharge against a higher back pressure. 
Double-tube injectors are used almost exclusively in locomotive work. 

281. Automatic Injectors. — Fig. 283 shows a section through the 
Penberthy injector. Its operation is as follows: Steam enters at the 
top connection and blows through suction tube c into the combining 
tube d and into chamber g, from which it passes through overflow valve n 
to the overflow m. When water is drawn in from the suction intake 
and begins to discharge at the overflow, the resulting condensation 
of the steam creates a partial vacuum above the movable ring h and 
the latter is forced against the end of tube c, cutting off the direct flow 
of water to the overflow. The water then passes into the boiler. Spill 
holes i, i, i are for the purpose of relieving the excess of water until 
communication with the boiler has been established. The action of 
opening and closing the overflow is entirely automatic. Where the 
conditions are not too extreme the automatic injector is to be preferred 
for stationary work because of its restarting features. It is also used on 
traction, logging, and road engines, where its certainty of action and 
special adaptability render it invaluable for the rough work to which 
such machines are subjected. 

Injectors, Theory of: Trans. A.S.M.E., 10-339; Sibley Jour., Dec, 1897, p. 101; 
Power, May, 1901, p. 23; Thermodynamics of the Steam Engine, Peabody, Chap. 
IX; Theory of the Steam Injector, Kneass. 

Injectors, General Description: Engr. U.S., Oct. 1, 1907, Nov. 15, 1907, July 15, 
1904, p. 501, Feb. 2, 1903, p. 151; Power, Aug., 1906, p. 478; Engr., Lond., March 
10, 1905, p. 244; Engineering, Aug. 30, 1895, p. 281. 



548 



STEAM POWER PLANT ENGINEERING 



w = 



(136) 



383. Performance of Injectors. — The performance of an injector 
may be very closely determined from the equation 

xr + q - t + 32 (Kneass, " Theory of the 
t-t Injector/' p. 83). 

in which 

w = pounds of water delivered per pound of steam supplied. 

x = quality of the steam supplied. 

r = heat of vaporization. 

q = heat of the liquid. 

t — temperature of the discharge water. 

t = temperature of the suction water. 

Figs. 284a, 284b, and 284c give the performance of a Desmond auto- 
matic injector as tested at the Armour Institute of Technology. The 
results check very closely with those calculated from above equation. 
Referring to Fig. 284a it will be seen that the weight of water delivered 
per pound of steam decreases as the initial pressure is increased, all 
other factors remaining the same. From Fig. 284b it will be noted 
that the weight of water delivered per pound of steam decreases as the 
temperature of suction supply is increased up to a point where the 
injector " breaks " or becomes inoperative. This critical temperature 
varies with the different types of injectors, being highest for the 
double-tube type, but seldom exceeds 160 degrees F. Fig. 284c shows 
that the weight of water delivered per pound of steam is practically 
constant for all discharge pressures within the limits of the apparatus. 

Table 71 gives the range of working steam pressures for standard 
" Metropolitan " injectors with varying suction heads and temper- 
atures, and, though strictly applicable to this particular type only, 
is characteristic of all makes. 

In selecting an injector the following information is desirable for 
best results: 

1. The lowest and highest steam pressure carried. 

2. The temperature of the water supply. 

3. The source of water supply, whether the injector is used as a 
lifter or non-lifter. 

4. The general service, such as character of the water used, 
whether the injector is subject to severe jars, etc. 

Injectors, Tests of: Eng. News, March 17, 1898, July 16, 1896, p. 39; Locomotive 
Engineering, May, 1900, p. 204; Power, Oct., 1904, p. 602; Railroad Gazette, Dec. 
11, 1896; Thermodynamics of the Steam Engine, Peabody, Chapter IX; Theory of 
the Injector, Kneass. 






PUMPS 



549 



SA 



19 



18 



IS 

p m. r 



15 



13 







































Constant Discharge Pressure 

20 Lb. Per Sq.In. 

Constant Suction Temp. 

55 Deg.Fah. 






























































CN 
































v () 







































































65 



70 



75 



90 



95 



Initial Gauge Pressure,Lb.Per Sq.In. 
Fig. 284a. Performance of an Automatic Injector with Varying Initial Pressure. 



20 

u 19 

la 18 

> « 

^£ 17 

P<3 17 

|^« 16 

3 15 

U 








































Constant Initial Pressure 

70 Lb. Per Sq. In. 

Constant Discharge Pressure 

70 Lb. Per Sq. In. 








( 


)^ 






















( 


5*^* 










































C 

































55 



65 75 85 95 105 

Temperature of Suction, Deg. Pah. 



115 



Fig. 284b. Performance of an Automatic Injector with Varying Suction Temperature 



■a 

£ a 
p<« 

u ° 

2 * 



20 



19 



18 



a* 



3 17 



16 





Constant Initial Pressure,70 Lb. Per Sq. In. 
Constant Suction Temperature, 56 Deg. Pah. 












c 


) 








- ( 


) 






















1 


y 










1 


i 


( 


J 

































30 AO 50 60 70 

Discharge: Pressure, Lb. PecSq. In. Gauge 



Fig. 284c. Performance of an Automatic Injector with Varying Discharge Pressure. 



550 



STEAM POWER PLANT ENGINEERING 





TABLE 71. 

RANGE IN WORKING PRESSURES. 

Standard " Metropolitan " Steam Injectors. 






Automatic. 


Suction 
Temperature, 




Suction Head, Feet. 




Degrees F. 


2 


8 


14 


20 


Under 
Pressure. 


Under 60 


25 to 150 


30 to 130 


42 to 110 


55 to 85 


20 to 160 


100 


26 to 120 


33 to 100 


55 to 80 




25 to 125 








120 










26 to 85 














140 


























Double Tube. 


Suction 
Temperature, 


Suction Head, Feet. 


Degrees F. 


2 


8 


14 


20 


Under 
Pressure. 


Under 60 


14 to 250 


23 to 220 


27 to 175 


42 to 135 


14 to 250 


100 


15 to 210 


26 to 160 


37 to 120 


46 to 70 


15 to 210 


120 


20 to 185 


30 to 120 


42 to 75 




20 to 185 








140 


20 to 120 


35 to 70 






20 to 120 











283. Injector vs. Steam Pump as a Boiler Feeder. — From a purely 
thermodynamic standpoint the efficiency of an injector is nearly 
perfect, since the heat drawn from the boiler is returned to the boiler 
again, less a slight radiation loss. As a pump, however, the injector is 
very inefficient and requires more fuel for its operation than very 
wasteful feed pumps. This is best illustrated by an example: An 
injector of modern construction will deliver say 15 pounds of water to 
the boiler per pound of steam supplied, with delivery temperature of 
150 degrees F. This corresponds to a heat consumption of 71 B.T.U. 
per pound of water delivered, thus : 

With initial pressure of 115 pounds absolute, 

X = 1185. 



PUMPS 551 

Heat in the water delivered to the boiler, 

150 - 32 = 118 B.T.U. above 32 degrees F. 
Heat of 1 pound of steam above a feed temperature of 150 degrees F. 

1185 -118 = 1067 B.T.U. 
Heat required to deliver 1 pound of water to the boiler, 
1067 



15 



71 B.T.U. 



A simple direct-acting duplex pump consumes say 200 pounds steam 
per I.H.P. hour. Assume the extreme case where the exhaust steam 
will not be used for heating the feed water and the latter is fed into 
the boiler at 60 degrees F. 

The heat supplied to the pump per I.H.P. hour, 

200 J1185 - (60 - 32)} = 231,400 B.T.U. 

Assuming the low mechanical efficiency of 50 per cent, the heat 
required to develop one horse power at the water end will be 

231,400 + 0.50 = 462,800 B.T.U. per hour. 

Since the steam pressure is 100 pounds gauge, the equivalent head 
of water at 60 degrees F. is 

2.3 X 100 = 230 feet. 

Assume the friction in the feed pipe, the resistance of valves, etc., to 
be 30 per cent of the boiler pressure; the total head pumped against 
will be 

230 + 69 = 299, say 300 feet. 
1 horse-power hour = 1,980,000 foot-pounds per hour. 
1,980,000 



300 



= 6600 pounds, 



that is, 1 horse power at the pump will deliver 6600 pounds of water per 
hour to the boiler against a head of 300 feet. 

The heat consumption per pound of water delivered, 

<f» = 7<UB.T.T,. 

If the feed water is heated to say 210 degrees F. by the exhaust steam 
from the pump, the heat consumption will be 55.5 B.T.U. as against 70.1 
without the heater. 

Thus even in this extreme case of poor steam-pump performance 
the heat consumption lies in favor of the pump. With the better 



552 



STEAM POWER PLANT ENGINEERING 



grades of pumps this disparity is considerably greater, and decidedly 
so if the exhaust steam is used to preheat the feed water. For inter- 
mittent operation the condensation losses in the pump may more 
than offset this gain. Other conditions, however, such as compact- 
ness, low first cost, and ease of operation are oftentimes considerations, 
and the heat consumption is of minor importance. 

284. Air Pumps. — Condenser air pumps may be divided into two 
classes : 

1. Wet-air pumps and 

2. Dry-air pumps. 

The former handle both air and water and the latter air alone. 
Ordinary jet-condenser wet-air pumps handle simultaneously the 
circulating water, condensed steam, and entrained air, and are, in 
fact, a combination of circulating and vacuum pump. Surface- 
condenser wet-air pumps are the same in principle and design, but 
are smaller in size for a given main engine output, as they handle the 
condensed steam and air only. 

Wet-air pumps may be driven by the main engine or independently 
and may be direct acting, Fig. 203, or fly wheel driven, Fig. 228. 
The fly-wheel type may be steam, electric, or belt driven. Dry-air 

pumps are virtually air compres- 
sors, as their function is to com- 
press air from the pressure existing 
in the condenser to that of the 
atmosphere. They are generally 
of the fly-wheel type. Where a 
high degree of vacuum is neces- 
sary the air cylinders are com- 
pounded, as efficient compression 
of say \ pound to 15 pounds or 30 
to 1 is too great for a single stage. 
385. Dean Air Pump. — Fig. 285 
shows a section of the air cylinder 
of a Dean twin-cylinder wet-air 
pump as applied to a jet condenser. 
There are three sets of valves, the 
suction or foot valves A, A, the lift- 
ing or bucket valves B, B, and the 
head or discharge valves C, C. On 
the upward stroke of the piston or bucket a partial vacuum is formed in 
the chamber between the bucket and the lower head, causing the water 




PUMPS 553 

and air in the bottom of the barrel to lift the foot valves A, A from 
their seats and flow into the cylinder. On the downward stroke the foot 
valves A, A close and water and air are entrapped in chamber R between 
the lower head and the bucket. As the bucket descends, the pressure 
of air in the cylinder lifts the bucket valves B, B from their seats and 
permits the air and water to escape to the upper portion S of the 
cylinder between the head plate and the bucket. On the next upward 
stroke the water and air are forced through the discharge valves C, C 
into the hot well. This discharge of water and air from the top com- 
partment is simultaneous with influx of water and air in the lower 
chamber. 

See paragraph 209 for other types of wet-air pumps in connection 
with jet condensers. 

286. Size of Wet-Air Pumps ; Jet Condensers. — In proportioning 
such pumps the quantity of cooling water and condensed steam to be 
taken care of is readily determined, but the percentage of air mingled 
with it must be estimated. Surface water under atmospheric pressure 
ordinarily contains from 2 to 5 per cent of air by volume. To pro- 
vide for possible leakage a very liberal factor is usually allowed, an 
average figure being about 10 per cent. 

Let Q = total volume of air and water in cubic feet per hour to 
be handled by the pump. 
V = volume of cooling water in cubic feet per hour. 
v = volume of condensed steam in cubic feet per hour. 
v a = volume of air at pressure P a and temperature T a . 
V + v = total volume of water and condensed steam at atmospheric 
temperature. 
T a = temperature of the air entering the condenser, degrees F. 
T 2 = temperature of the discharge water, degrees F. 
T = initial temperature of the cooling water, degrees F. 
P a = atmospheric pressure, pounds per square inch. 
P c = total pressure in the condenser, pounds per square inch. 
P v = pressure of aqueous vapor at temperature T 2 . 

Then (V + v) = volume of water to be pumped from the condenser 
per hour and v a == volume of air at atmospheric pressure and tempera- 
ture T a entering the condenser; but on entering the condenser the air is 
increased in volume, due to the reduction in pressure and the increase 
in temperature, and the total volume to be exhausted per hour by the 
pump is 

P T + 460 

Q=V + v + VaJ J^ T ^ t ± r ^ Q . (13 8) 



554 



STEAM POWER PLANT ENGINEERING 



Under average conditions of reciprocating engine practice v = 5 V V, 
P a = 15 pounds per square inch, P c = 2 pounds per square inch, 
P v = 1.27, T 2 = 110 degrees F., T Q = 60 degrees F. Substituting 
these values in above equations, assuming v a = 10% of V, we get 
Q = 3.4 V. 

Average practice gives 3 V as the pump displacement per hour for 
a single-acting pump and 3.5 V for a double-acting pump, the cylinders 
being ordinarily proportioned on a basis of 50 feet per minute piston 
velocity at rated capacity. 

Table 72 gives the approximate sizes of air pumps for condensers as 
manufactured by prominent makers. 

TABLE 72. 

APPROXIMATE SIZES OF AIR PUMPS FOR CONDENSERS. 



Pounds of 
Steam Con- 
densed per 
Hour. 


Jet Condenser. 


Surface Condensers. 


Duplex 
Pump. 


Horizontal 

Double Acting 

Pump. 


Vertical 

2-Cylinder 

Single 

Acting. 


Horizontal. 


Vertical 
2-Cylinder. 


500 to 1,000 
1,000 to 1,500 


4|X 5 
5f X 6 
6| X 6 
7iX 6 

7 X 10 

8 X 10 
8£X 10 

9 X 10 
10 X 10 
10* X 10 

11 X 10 

12 X 10 
12 X 15 
15 X 15 
17 X 15 
19 X 15 


6X 7 
8X 7 
8X 12 
9X 9 
9X 10 

11 X 12 

12 X 14 
14 X 14 

14 X 16 

15 X 16 

15 X 18 

16 X 18 
18 X 18 
20 X 24 
24 X 24 
26 X 24 


5X 4 
6X 4 
7X 5 
9X 6 

10 X 8 

11 X 9 

12 X 8 
12 X 10 

14 X 10 

15 X 10 

15 X 12 

16 X 10 

17 X 12 
20 X 12 
22 X 15 
24 X 18 


3*X 4 
4X4 
4X6 
5X7 
5X8 
6X8 
7X9 
7 X 10 

7 X 12 

8 X 10 

8 X 12 

9 X 12 
10 X 12 
12 X 14 
14 X 16 
16 X 24 






1,500 to 2,000 




2,000 to 2,500 




2,500 to 3,000 




3,500 to 4,000 




4,000 to 4,500 




4,500 to 5,000 




5,000 to 6,000 




6,000 to 7,000 

7,000 to 8,000 

8,000 to 9,000 

9,000 to 10,000 

10,000 to 15,000 

15,000 to 20,000 

20,000 to 25,000 


8X 4 
8X 6 
9X 6 
10 X 8 
11X 8 
12 X 8 
14 X 10 



Wet-air pumps are usually independently driven, making it possible 
to vary the speed of the pump irrespective of the engine speed and to 
create a vacuum before starting the engine. Occasionally, however, 
when the load is constant, as in pumping-engine practice, the pump 
may be driven by the main engine. 

Air Pumps: Power, Nov., 1904, p. 652; Engineering, Nov. 17, 1905; Engr. U.S., 
Jan. 1, 1906, Elec. Engr., Lond., Nov. 14, 1902; American Mach., Jan. 31, 1901, 
p. 113; National Engr., June, 1906. 

Air Pumps. — Reference Books: Hausbrand, Evaporating and Condensing 
Apparatus, p. 383; Whitham, The Steam Engine, p. 301; Thurston, Manual of the 
Steam Engine, Vol. II, p. 145; Seaton, Manual of the Marine Engine, p. 301. 



PUMPS 



555 



287- Edwards Air Pump. — Fig. 286 shows a section through the 
air cylinder of an Edwards air pump. This device belongs to the sur- 
face-condenser " wet-air pump " class, as both the water of condensa- 
tion and the entrained air are exhausted simultaneously by the same 
piston. Unlike the standard 
type of wet-air pumps, foot 
valves and bucket valves are 
entirely dispensed with. The 
condensed steam flows contin- 
uously by gravity from the 
condenser into the base of the 
pump through passage A and 
annular space B. As the piston 
C descends it forces the water 
from the lower part of the 
casing F into the cylinder 
proper through the ports P, P. 
On the upward stroke the 
ports in the piston are closed 
and the air and water dis- 
charged through head valves 
D and exhaust port E to 
the hot well. The seats of 




Fig. 286. Edwards Air Pump. 



valves D are constructed with a rib between each valve and a lip 
around the outer edge, so that each valve is water-sealed independently 
of the others. In earlier air pumps of this general type the clearance 
between the bucket and head valve seat is necessarily large, due to 
the space occupied by the bucket valves and the ribs on the under side 
of the valve seating. This clearance space reduces the capacity of the 
pump, since the air above the bucket must be compressed above 
atmospheric pressure before it can be discharged, and on the return 
stroke will expand and occupy a space which should be available for a 
fresh supply of air from the condenser. In the Edwards air pump the 
clearance space is reduced to a minimum, since there are no bucket 
valves to limit it. The absence of suction or foot valves still further 
increases the capacity of the pump for similar reasons. These pumps 
are arranged either single, double, or triplex; steam, electric, or belt 
driven; slow or high speed. They are ordinarily used in connection 
with surface condensers. 

Edwards Air Pump: Engr., July 1, 1903, p. 536; Engineering, 62-221, 63-60, 
64-767, 80-328; Eng. News, June 12, 1902, p. 478. 

Centrifugal Wet Vacuum Pump: Power & Engr., Jan. 4, 1910. 



556 



STEAM POWER PLANT ENGINEERING 



288. Mullan Valveless Air Pump. — Fig. 287 gives several views of 
the " Mullan valveless air pump " as used in connection with the 




Half Section on 
Line C-D 
Sad Elevation 
to Line £-E 



Elevation on 
Line A-B 

Half Section on 
C.L. oX Air Cylinder 



Fig. 287. Mullan Valveless Dry-Air Pump. 



C. H. Wheeler Company's " high-vacuum " condensing outfit. The 
pump is double acting and devoid of suction valves. The cylinder 



EXHAUST 



CONDENSER 




MAIN PUMP DRY AIR PUMP 

Fig. 288. Hewes and Phillips Air Pumps. 



has a central port which is uncovered by the piston at each end of 
the stroke and covered at all other positions. Discharge valves of 



PUMPS 



557 



the familiar spring-seated poppet type are located in both heads of the 
cylinder. As the piston moves from one end of its stroke to the 
other it forms a vacuum behind it and forces out the gases and water 
ahead of it; when it reaches the end of the stroke the central inlet 
port is uncovered and the vacuum behind the piston draws in the 
condensation and gases from the condenser. This operation is repeated 
on the return stroke. 

The makers claim that the pump will operate, under shop test 
conditions, within one-half inch of the barometer, enabling them to 
guarantee a vacuum within two inches of absolute under full-load 
conditions of steam turbine operation. 

289. Alberger Rotative Dry-Air Pump. — Fig. 289 shows a section 
through the air cylinder of an Alberger rotative dry-air pump, illus- 




ZZZZ W& 



Fig. 289. Alberger Rotative Dry- Air Pump. 

trating a type of pump in which the admission valve is mechanically 
operated. This pump is designed to operate with dry air only, all 
condensation being removed before the air enters the cylinder. This 
permits of the use of a small clearance space and makes it possible to 
run at a higher speed of rotation than can be secured with a type of 
pump in which water is used to seal the valves. Referring to Fig. 289, 
air is being taken into the right-hand end of the cylinder through inlet A 
and forced from the left-hand end through exhaust opening B. Rotary 
valve mechanically opens to admission and mechanically closes the 
discharge. The discharge opening depends on the spring-regulated valve 
C at the top of the cylinder. Heads are water jacketed. Ports and 
passages are made large to reduce the friction of the air entering the 
pump, and to obviate the bad effects of clearance, an equalizing passage 



558 STEAM POWER PLANT ENGINEERING 

is provided in valve 0. The action of the passage is shown in the 
section to the right. When the piston reaches the end of the stroke 
the clearance space is filled with air at atmospheric pressure. If this 
pressure were not relieved the piston would travel a considerable dis- 
tance before drawing in air from the condenser. By means of the 
equalizing passage the clearance space is connected to the opposite end 
of the cylinder and the vacuum there reduces the pressure in the clear- 
ance space. 

290. Size of Wet-Air Pump for Surface Condenser. — Since the wet- 
air pump handles both the air and condensed steam, its theoretical 
capacity, neglecting clearance, may be determined by eliminating V 
from equation (138), which then becomes 

For the average reciprocating engine, P c = 2, T 2 = T a = 110, 
T = 60, P v = 1.27. Assuming P a = 15 pounds per square inch, and 
substituting these values in (139), Q = v + 25 v a . 

The volume of air entering the condenser varies so much with the 
character of the power plant equipment and the conditions of operation 
that any assumed " average " value of v a may lead to serious error. 
(See Power & Engr., Feb. 2, 1909, p. 234.) 

A study of some two hundred condenser installations gives 

Q = 10 v for the average reciprocating engine and 
Q = 20 v for average steam turbine practice. 

Table 72 gives the cylinder dimensions of wet-air pumps as advo- 
cated by prominent condenser builders. 

291. Size of Dry- Air Pumps. — " Dry -air " pumps are used in con- 
nection with barometric and surface condensers where a high degree 
of vacuum is essential, as in steam turbine practice. Such pumps are 
intended to exhaust the saturated non-condensable vapors only. 

The capacity of the dry-air pump is based upon experience rather 
than theory. An investigation of some fifty installations gives 

Q = 20 v to 30 v for vacua under 27 inches. 

Q = 50 v for vacua of 28 inches and over, both referred to a 30-inch 
barometer. 

Professor Weighton states that " with suitable condenser arrange- 
ments and a reasonably air-tight system there is nothing gained in 
efficiency by the use of air pumps exceeding in capacity 0.7 of a cubic 
foot per pound of steam condensed up to a vacuum of 29 inches. " 
(Engineering Record, May 19, 1906, p. 61.) 






PUMPS 559 

The work done by the average " high- vacuum " dry-air pump is a 
maximum for vacua between 18 and 20 inches. 

This may be proved from Fig. 290, which represents a theoretical 
indicator card from the air-pump cylinder. 

Let p 2 = pressure in the condenser, pounds per square inch absolute. 
p x = atmospheric pressure. 

v 2 = piston displacement, including clearance, cubic feet. 
v x = volume of air in the cylinder when the valve opens to 

atmosphere, cubic feet. 
v c = clearance volume, cubic feet. 

The work done is proportional to the area ABCD. 
Area ABCD = work done = area EBIO + BAGI -FAGO - ECDF. 
Neglecting the exponential factor n for the sake of simplicity, thus 
making pv = p x v x = p 2 v 2 = constant, 



W 



= work done = p x v x + p x v x I * — — p 2 v 2 — p x v e + p 2 v e . (140) 

Jv 2 V 



Substitute p x v x for its equivalent p 2 v 2 and - £ - i — 3 — for its equiva- 
lent p 2 v c and integrate. 

W = p x v x + p t v x log e v 2 - p x v x log, v x - p x v x + £l^, (141) 



making the first derivative zero. 




^=0=\og e v 2 -l-\o ge v x + -, 


(142) 


= log e - 2 -l +- c , 
v x v 2 


(143) 


i.e., W is a maximum when 




1 ^2 V C 

- log, -? = - c - 1 


(144) 


1 Pi -. ^C 

or loge J - L = 1 j since p.f, = 79 9 v . 

?2 ^2 


(145) 



For average high-vacuum practice v c = 3 per cent of the* piston dis- 
placement. Assume v 2 = 1, v c = 0.03, and p x = 14.7 pounds per square 
inch, and substitute these values in (145), thus: 

14.7 , 0.03 

l0ge ^7 =1 -m)- 

Whence p 2 = 5.5 pounds per square inch absolute, which corresponds 
to a vacuum of 18.6 inches of mercury. 



560 



STEAM POWER PLANT ENGINEERING 



Thus we see that the maximum load on the pump occurs when the 
vacuum is between 18 and 20 inches. If the vacuum is less than this, 
the load falls off because of the decreased difference in pressure. If the 




Fig. 290. 

vacuum is greater, the load falls from the decrease in weight of air 
handled. 

291a. Hot-well Pumps. — In high-vacuum surface condensers the 
condensed steam is often handled by a small independently driven pump 
called the hot-well pump. See paragraph 228. 

Types of Air Pumps: Power & Engr., June 1, 1909, p. 963. 

292. Centrifugal Pumps. — Centrifugal pumps consist of two essen- 
tial elements, (1) a rotating impeller which draws in the water at its 
center, and (2) a stationary casting which guides the water thrown 





Fig. 291. Types of Impellers. 

from the ends of the impeller to the discharge outlet. The impellers 
may be of the open type, Fig. 291(B), or closed, Fig. 291(A). The casing 
may be cylindrical and concentric with the impeller, Fig. 295, or of 






PUMPS 



561 



spiral form, Fig. 292. The shape of the impeller and casing determines 
the efficiency of the pump and its adaptability to certain conditions of 
service. 

Centrifugal pumps may be classified as 

(1) Volute. 

(2) Turbine. 

293. Volute Pumps. — Fig. 292 shows a section through a typical 
volute pump. The casing is of spiral design, forming a gradually 

increasing water or " whirlpool " chamber 
for the purpose of partially converting 
velocity head to pressure head. The older 
forms of volute pumps were very ineffi- 
cient, seldom delivering more than 50 per 
cent of the energy supplied, and usually 
not adapted to lifts greater than 50 feet. 
The modern pump gives efficiencies as 
high as 77 per cent, and lift is limited only 
by the speed of the impeller. As a general 
rule the volute pump is of single-stage 
construction, and best adapted to com- 
paratively low lifts (under 80 feet), though 
an exception is found in the De Laval 

steam turbine driven volute pumps, which are made both single and 

multi-stage for lifts as high as 700 feet. 

294. Turbine Pumps. — The directions of flow in the casing and from 
the impeller in a volute pump are at cross currents with each other, as 
shown in Fig. 293. The turbine pump, Fig. 294, is provided with a 




Fig. 292. 



A Typical Centrifugal 
Pump. 




Fig. 293. Direction of Water from the Impellers 
of a Centrifugal Pump without Diffusion Vanes. 



Fig. 294. Effect of Diffusion Vanes 
on the Direction of Water. 



system of diffusion vanes or expanding ducts disposed between the 
periphery of the impeller and the annular casing somewhat like the guide 
vanes in a reaction turbine water wheel, so that the fluid emerges 



562 



STEAM POWER PLANT ENGINEERING 



tangentially at about the velocity in the casing. The casing is con- 
centric with the impeller, since the diffusion vanes render the volute 




Fig. 295. Lea-Degan Three-Stage Turbine Pump. 

chamber unnecessary. For high lifts these pumps are compounded, 
thereby reducing the peripheral velocity and decreasing the friction 
losses. Fig. 295 shows a section through a three-stage turbine pump 
and Fig. 296 a section through a six-stage pump. 




Fig. 296. Rateau Six-Stage Turbine Pump. 

In view of past developments it is probable that the centrifugal pump 
will supplant to a considerable extent the present type of piston pump 



PUMPS 563 

for many uses. Efficiencies above 70 per cent are not unusual, and 
the head against which the pump may operate is limited only by the 
peripheral speed at which the impeller may be safely run. Some of 
the advantages of the modern high-grade centrifugal pump as compared 
with the piston type are: 

(1) Low first cost, 

(2) Compactness, 

(3) Absence of valves and pistons, 

(4) Low rate of depreciation, 

(5) Uniform pressure and flow of water, 

(6) Simplicity of design and ease of operation, 

(7) Freedom from shock, 

(8) High rotative speed, permitting direct connection to electric 

motors and steam turbines, 

(9) Ability to handle dirty water, sewage, and the like, 

(10) In case of stoppage of delivery, the pressure cannot increase 

beyond the predetermined working pressure, and 

(11) Ease of repair. 

Some of the disadvantages are: 

(1) Efficiency not as high as the best grade of piston pumps, 

(2) Cannot be direct connected to low-speed engines when high 

lifts are desired, and 

(3) The rate of flow cannot be efficiently regulated for wide ranges 

in duty. 

Theory of Centrifugal Pumps: Engr. U.S., Oct. 1, 1907, p. 908; Bulletin of Univ. 
of Wisconsin, No. 173; Trans. A.S.M.E., 54-470; Prac. Engr., Lond., Sept. 29, 1905; 
Jour. Asso. Eng. Soc, Dec, 1901; Centrifugal Pumps, Turbines, and Water Motors, 
C. H. Innes, Chap. XXVI. 

Centrifugal Pumps, General Description: Trans. A.S.M.E., 26-764; Engr. U.S., 
Aug. 1, 1907, p. 723, Oct., 1907, p. 908, Oct. 15, 1907, p. 952; Power, March, 1907, 
p. 172; Eng. Mag., July, 1906; Elec. World, Jan. 12, 1907, p. 113; Eng. Rec, Feb. 9, 
1907, p. 165; Machinery, Jan., 1907, p. 237, April, 1907, p. 442. 

295. Performance of Centrifugal Pumps. — For best efficiency a 
centrifugal pump must be properly designed for the intended service 
as to curvature of vanes, diameter and speed of impeller, and num- 
ber of stages. Figs. 298 to 300 are based upon experiments with 
De Laval centrifugal pumps. When a practically uniform head is 
required at constant speed with varying water supply as in city water- 
works, hydraulic elevator systems, or boiler feeding, the impeller vanes 
are designed to give the characteristic curve illustrated in Fig. 298, 
which protects the motor from possible overload. 



564 



STEAM POWER PLANT ENGINEERING 



In dry-dock and other variable-head work, in order not to overload 
the motor, the power should be practically constant through wide 
variations of head and at the same time the efficiency should not vary 
seriously. A desirable characteristic for such a pump is illustrated in 
Fig. 299. 



H 





















































Curves of Performance 

For 
Morris Centrifugal Pump 

Variable Speed 

Rated Cap. For 60 Ft. Hd. 

120 Gal. at 830 R.P.M. 




















50 






























































































































































/ 


40 








































/ 








































f 




























E 


ff. 






























^ 


^' 


7 


/ 






































/ 


/ 








































/ 










tzft 




























/ 










^Z_ 




t/ 
























/ 












4/ 




^r~ 




A 


































f 




* 


sy 
















{/ 






















f 


V 
















M 
































20 




/ 










































/ 










































/ 










































/ 
















































































10 














/ 
































/ 




y 


', 


y 






y 


y 








c 


^ 


j^> 














/ 


^ 


X 




































\ 


r 








































J 




== 


^ 




— - 


-»" ■ 





























200 



100 



200 



300 



400 500 

Revolutions Per Min. 



600 



700 



Fig. 297. 



In water-supply systems in which the friction of the piping is a large 
part of the total head at full delivery, the characteristic shown in Fig. 
300 is especially useful. Thus, when the system reduces its demand 
for water and the frictional head is consequently considerably reduced, 



160 












































































































































L 






















\ 








m 


— ■ 






















g 


J^- 










































120 












GU 








































































































































































































































80 






































































































^> 








































A\C 


V 




















\ 


































<m- 










































40 












^ 


^ 





































































































































































































































































































40 



140 



60 80 100 130 

Capacity 

Fig. 298. Centrifugal Pump Characteristic for Hydraulic Elevator Service, Boile* 

Feeding, etc. 



















































































































120 
















































































































100 - 
























































•o 8°- 




























Ef 


tici 


eirr 


V 


















1 " 

W Ail 


























c£ 


Itii 












-=> 














































*fc 


•6, 


p* 














































































































































































































































































































20 40 60 80 100 120 140 

Capacity 
Fig. 299. Centrifugal Pump Characteristic for Dry-Dock Service. 


























































160 
























































































































































































































































































^100 - 














Ch 


im 


?.t.e 


us 


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40 



100 120 



140 



60 80 

Capacity 
Fig. 300. Centrifugal Pump Characteristic for Water Works with Large Friction Head. 



566 



STEAM POWER PLANT ENGINEERING 



6 

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PUMPS 



567 



the pump would automatically adjust itself to the reduced head without 
change of speed. Figs. 297 to 304 are based upon experiment and show 
the relationship between speed, head, capacity, efficiency, and power 
consumption of various types of pumps. 

Tables 73 and 74 give the capacity, speed, head, and power require- 
ments for commercial sizes of centrifugal pumps, and may be used as a 
guide in selecting the size of pump for general service. 

TABLE 74. 
DATA PERTAINING TO WORTHINGTON MULTI-STAGE TURBINE PUMPS. 





c 
o 

1 oi 

•3 § 

is 

5 




— — 

P-l o 

11 

* 


Total Head in Feet, R.P.M., Number of Stages. 


Diameter of Di 
charge Pipe, 
Inches. 


100 Ft. 


200 Ft. 


300 Ft. 


400 Ft. 


S 
fc 

ti 


o . 

If 

2 
2 
2 
2 
2 
2 
1 
1 
1 
1 




o . 

& 1 


§ 
P^ 

tf 


if 

§00 


3 


t-r O) 


1 


1.5 

2 

2.5 

3 

4 

5 

6 

7 

10 
12 


30 

45-60 

75-100 

125-150 

200-250 

350-450 

600-700 

800-1000 

1500-1800 

2500-2800 


0.02 

0.0395 

0.0625 

0.095 

0.134 

0.222 

0.297 

0.396 

0.643 

1.00 


2000 

1500 

1300 

1200 

1100 

950 

800 

750 

600 

500 












1.5 


2000 
1800 
1600 
1400 
1200 
1300 
1200 
1000 
800 


3 
3 
3 
3 
3 
2 
2 
2 
2 










2 


1500 
1300 
1200 
1000 
1150 
1000 
800 
700 


4 
4 
4 
4 
3 
3 
3 
3 






2.5 
3 
4 
5 
6 
8 
10 


1050 
950 
780 
670 


5 
5 
5 
4 
4 
4 
4 



* Horse power based on maximum capacity. 



Tests of Centrifugal Pumps: Engr. U.S., Oct. 15, 1906, p. 685 ; Eng. and Min. Jour., 
April 14, 1906, p. 698; Eng. News, June 2, 1904, p. 512; Eng. Rec, July 1, 1905, 
p. 25, Sept. 29, 1906, p. 352; Iron Age, Sept. 1, 1904, p. 25; Machinery, Nov., 1906, 
p. 144; Power, Nov., 1906, p. 688; Trans. A.S.M.E., 22-262, 831; Jour. Am. Soc. 
Naval Engrs., 17-85. 

296. Rotary Pumps. — Rotary pumps are often used for circulating 
cooling water in condenser installations, and give about the same 
efficiency as centrifugal pumps under similar conditions of operation. 
For moderate pressure and large volumes they offer the advantage of 
low rotative speed, thus permitting direct connection to slow-speed 
steam engines. At high speeds they are noisy, due chiefly to the 
gearing. They occupy considerably less space than piston pumps of 
the same capacity, but require more room than the centrifugal type. 

Fig. 305 shows a section through a two-lobe cycloidal pump. The 
shafts are connected by wheel gearing, the power being applied to one 
of the shafts. The water is drawn in at / and forced out at 0, the 



568 



STEAM POWER PLANT ENGINEERING 



|50 
45 
40 
35 
30 
25 
20 
15 
10 
5 

1 




















1 








































































































































































































., 
















7" 




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H 


















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V 




































/ 




































/ 


<' 






































/ 






































/ 






































/ 






































7 


/ 


1 




































/ 






































/ 






































/ 









































«o 



800 900 1000 



100 200 300 400 500 000 700 

Capacity, Gallons per Minute. 

Fig. 301. Performance of a Six-Inch Worthington Conoidal Pump 




TOO 200 300 400 600 



Fig. 302. Performance of a Single-Stage De Laval Centrifugal Pump. 



PUMPS 



569 















































oo 












































to 












































no 


* 










































■OO 


i 










































no 


Z 
















































TOTAL LirT. rEET 














































8 






1 
u 














«rp 


^ejjS 


1— 9- 
























70 


°-~^. 








60 


< 

K 












*2 




















\ 
























*£j 


Z*~ •" 






~~~ 






^V 


N 
















b 










05 




























W 


/ 








































?0 


R 


/ 








































10 


z 


s' 










































r 
















CALL 


DNS f 


•ER N 


1INUT 


C 


















J 


zoo 400 eoo eoo iooo isoo 1400 1600 leoo zooo 2200 2400 2soo zeoo 3000 3200 3400 3600 3aoo 4Q00 



Fia. 303. Performance of a Two-Stage Lea-Degan Turbine Pump. 




50 100 ISO 200 J SO 300 )S0 400 

Fig. 304. Performance of a Two-Stage De Laval Centrifugal Pump. 



570 



STEAM POWER PLANT ENGINEERING 





Fig. 305. Two-Lobe Cycloidal 
Pump. 



Fig. 306. Rotary Pump with 
Movable Butment. 



3 65 



I 60£ 



3 * 

"3 55 



50 0.2, 
























Efiicien 


cy 


^y^ 








^ 






^^ 






^" 






,/ 


s 








t^ - ^^ 




/ 








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— ° 


/ 








^2 


^e^^-- 










O^ 
















Average I 


lead, 75 Ft. 



















300 



400 



500 



600 700 

Revolutions Per Minute 



900 



50 



45 



40 



30- 
o 

25=3 



20 



15 



1000 



£70 



2.4 



2.2 



o 2.0 

C4 



1 1.8 



I W 



1.0 



70 



Head Constant, Speed Variable. 

























1 




















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****' 


































. 








*# 


r^>- 




























jcjj2 


io^ 





















p$> 




Average Speed 812 R.P.M. 
Average Capacity 
45.5 Gal.Per Minute 


•^ 







































100 110 120 130 140 150 
Total Head, Eeet, 

Speed Constant, Head Variable. 



160 170 180 



Fig. 307. Performance of a Small Rotary Pump. 



PUMPS 



571 



displacement per revolution being equal to four times the volume of 
chamber A. There is no rubbing between impellers and casing. In 
this type of pump the 
pressure is independent 
of the speed of rotation, 
and the capacity varies 
almost directly with the 
speed. The slip varies 
from 5 to 20 per cent 
according to the dis- 
charge pressure. 

Fig. 306 shows a sec- 
tion through a rotary 
pump with movable but- 
ment. Fig. 307 illus- 
trates the performance 
of a 45-mm. Siemens- 
Schuckert rotary pump 
at different speeds and 
discharge pressures. 
(Zeit. d. Ver. Deut. Ing., 
June 24, 1905, p. 1040.) 
Large rotary pumps give 
much higher efficiencies, 
but the general charac- 
teristics are about the 
same. A combined effi- 
ciency of pump and 
engine as high as 84 per 
cent has been recorded. 
(Trans. A.S.M.E., Vol. 
24, p. 385.) 

Rotary Pumps: Engr. 
U.S., Jan. 1, 1904, p. 51 
Am. Mach., March 12, 1896 
p. 238, Jan. 25, 1906, p. 103 
Trans. A.S.M.E., 24-385 
Eng. News, -March 1, 1900 
p. 152; Power, Dec, 1903 
p. 59, Aug., 1905, p. 477; Fig. 308. 10,000,000 Gallon Circulating Pump. 

The Constructor, Releaux, p. 226; Eng. Rec, Jan. 10, 1903, p. 59. 

Tests of Rotary Pumps: Zeit. d. Ver. Deut. Ing., June 24, 1905, p. 1040; Trans. 
A.S.M.E., 28-503. 




572 



STEAM POWER PLANT ENGINEERING 



297. Circulating Pumps. — This term is ordinarily applied to the 
pumps which supply injection water to the condenser. 

These types are found in practice: the piston, the centrifugal, and 
the rotary pump. Figs. 224 and 236 show the application of recipro- 
cating pumps to condenser installations and Fig. 211a and Fig. 443 
a similar application of centrifugal pumps. 

For large volumes of water and low heads the centrifugal or rotary 

pump is generally adopted on ac- 
count of minimum space require- 
ments and low first cost. 

In very large central stations 
where the demand for circulating 
water is enormous and the lift is 
moderately high, the high-duty 
pumping engine is often installed. 
Fig. 308 shows a section through 
one of the nine high-duty circulat- 
ing pumps at the New York Rapid 
Transit Company's power house. 
The steam end is operated by 
Corliss cylinders and is of the 
cross compound type. The maxi- 
mum capacity is 10,000,000 gallons 
per day (24 hours) against a head 
of 50 feet at mean low water. The 
actual lift is much less than this, 
as the discharge is aided by the 
vacuum in the condenser. 

Pumping Engines: Engr. U.S., Dec. 1, 
1905, p. 786; Engng., Jan. 27, 1905, p. 132; 
Eng. Mag., Jan., 1905, p- 77 ; Eng. Rec, July 
8, 1905, p. 58; Eng. News, Dec. 6, 1905, 
May 26, 1904; Trans. A.S.M.E., 3-141, 
9-476, 12, 534, 975, 13, 83, 176, 14-1340, 
Fig. 309. The Pulsometer. 15-1103, 16, 49, 169, 21, 327, 788, 1018. 

298. Air Lift. — The air lift is a simple arrangement of piping 
whereby water may be raised by means of compressed air. There are 
no working parts, and no valves are employed except to regulate the 
supply of air. Its particular field of application lies in pumping water 
from a number of scattered wells, and on account of the total absence 
of working parts it is peculiarly adapted to handling water containing 
sand, grit, and the like The device consists of a partially submerged 







PUMPS 



573 



water pipe and an air-supply pipe variously arranged as in Fig. 310(A) 
to 310(D). Compressed air forced into the water pipe at or near the 
bottom forms a series of bubbles or " pistons," as shown in D, which 
displace an equal volume of water. (For the theory of the air lift see 
Compressed Air, October, 1905, p. 3696.) The pressure required to 
operate the lift after it is once started is considerably less than the 




WATER LEVEL_ 
WELL 



Fig. 310. Different Arrangements of the " Air Lift. 



pressure of water due to the head, while that required to start it is 

slightly greater, consequently the pressure may be reduced after the 

pump begins to work properly. The successful operation of this 

device depends upon the ratio of the depth of submersion A, Fig. 

310(D), to the total lift B, and the ratio of the area of the air pipe to 

that of the water pipe. The best results are obtained, in a general 

A 
sense, when the ratio — lies between 0.55 and 0.85, an average figure 

being 0.65, and when the area of the air pipe is 0.16 that of the water 
pipe. 

The quantity of air needed may be closely approximated by the 
following equation {Engineer, London, Aug. 14, 1903, p. 173): 



V = 



QL 
20 



in which 

V = cubic feet of free air per minute. 

Q = cubic feet of water per minute. 

L = lift in feet above the surface of the water. 

The velocity of the air should not exceed 4000 feet per minute. 



574 



STEAM POWER PLANT ENGINEERING 



The efficiency ("water" horse power divided by "air" horse 

power) varies from 30 to 50 per cent, increasing as the ratio — 

B 

increases from 0.55 to 0.85. {Engineer, U.S., Aug. 15, 1904, p. 564.) 

A number of tests give efficiencies (" water " horse power divided by 

I.H.P. of steam cylinder) varying from 20 to 40 per cent. The horse 

power required to compress one cubic foot of free air to different 

pressures per square inch, as determined from actual practice, is 

approximately as follows: 



Pressure in 
Pounds. 


Horse Power 
Required to 

Compress 
1 Cubic Foot. 


Pressure in 
Pounds. 


Horse Power 
Required to 

Compress 
1 Cubic Foot. 


176 
140 
100 

80 


0.434 
0.376 
0.201 
0.189 


60 
45 
30 


0.159 
0.145 
0.121 



(Engr., Lond., Aug. 14, 1903, p. 174, Dec. 11, 1903, p. 568, Feb. 12, 1904, p. 172.) 

When it becomes necessary to raise water to a height exceeding 
say 175 feet above the level in the well, it is customary to use two or 
more pumps, the total lift being divided between them. 

Air Lift: Engr., Lond., Aug. 14, 1903, p. 173, Dec. 11, 1903, p. 568, Feb. 12, 1904, 
p. 172, Jan. 10, 1908; Eng. Rec, Jan. 7, 1905, p. 8; Engineering, Jan. 22, 1904, 
p. 135, Jan. 29, 1904, p. 166; Compressed Air, Aug., 1903, Oct., 1905; Engr. U.S., July 
15, 1903, p. 547, Sept. 21, 1906, Aug. 15, 1904, p. 564; Mech. Engr., Aug. 20, 1904; 
Engng. Rev., July, 1904; Power, March, 1905, p. 173; Eng. News, Jan. 16, 1908. 

Pulsometer: Tech. Quar., Sept., 1901; Public Works, Aug. 15, 1904; Engr. U.S., 
July 15, 1904; Experimental Eng., Carpenter, p. 621; Thermodynamics, Wood, 
p. 293; Trans. A.S.M.E., 13-211. 

Cost of Operating American Pumping Stations: Eng. Rec, Aug. 6, 1904; Proc. 
Engrs. Club of Phil., Oct., 1906. 

Complete Description of Various American Types of Steam, Rotary, and Centrifugal 
Pumps: Engr. U.S., Jan. 1, 1904. 

The Growth of the Pumping Station: Eng. Rec, July 14, 1906, p. 50. 

The Selection of Waterworks Pumping Machinery: Eng. News, Vol. 52, p. 39. 

Centrifugal Pump for Boiler Feeding: Power & Engr., Mar. 15, 1910. 

Recent Records of High Duty Pumping Engine: Eng. News, Feb. 3, 1910. 



CHAPTER XIV. 

SEPARATORS, TRAPS, DRAINS. 

299. Live-Steam Separators. General. — The function of a steam 
separator is the removal of entrained water from steam. 

Unless a boiler is liberally provided with superheating surface, the 
steam may contain an amount of moisture varying from 0.3 to. 5 per 
cent. If the boiler is poorly proportioned or forced far above its rating, 
this percentage may be greatly increased. The quality of the steam 
is still further reduced by condensation in the steam pipe, which may 
vary from one to ten per cent, depending upon the length of pipe and 
efficiency of covering. 

One of the effects of moisture in steam is to increase its density 
and reduce its elastic- force. It also increases its conductivity, so that 
during the work of expansion more heat is absorbed from the walls of 
the cylinder and discharged into the atmosphere or into the condenser 
without doing useful work. (Ewing, " The Steam Engine," p. 151.) 
Although the heat loss from this cause is small, the danger arising 
from the introduction of a considerable amount of water in the 
cylinder renders the removal of the moisture necessary. See page 248 
for influence of moisture on steam consumption. 

The essentials of a good separator are high efficiency as a water 
eliminator, ample storage capacity for any sudden influx of water, 
simplicity and durability in construction, and small resistance to the 
current of steam passing through. A good separator may be relied 
upon to remove practically all of the moisture from steam containing 
under ten per cent entrainment and all but two per cent from steam 
containing as much as twenty per cent. (Engineer, U.S., Jan. 15, 
1904.) 

Table 75 gives the results of a series of tests made by Professor 
R. C. Carpenter in 1891 of six steam separators. (Power, July, 1891, 
p. 9.) Conclusions from these tests were: 

1. That no relation existed between the volume of the several 
separators and their efficiency. 

2. No marked decrease in pressure was shown by any of the separa- 
tors, the most being 1.7 pounds by separator E. 

575 



576 



STEAM POWER PLANT ENGINEERING 



3. Although changed direction, reduced velocity, and perhaps centri- 
fugal force are necessary for good separation, still some means must 
be provided to lead the water out of the current of the steam. 

A series of tests made at Armour Institute of Technology in 1905 on 
a number of separators showed that the efficiency of separation decreased 
as the velocity of the steam increased.* At the low velocity of 500 feet 
per minute all separators were equally efficient, at a velocity of 5000 
feet per minute several had little effect on eliminating the moisture 
present, and at a velocity of 8000 feet per minute only one gave efficient 
results. 



TABLE 75. 
TESTS OF STEAM SEPARATORS. 

(R. C. Carpenter.) 





Test with Steam of about 10 
per Cent of Moisture. 


Make of 
Separator. 


Quality of 
Steam 
Before. 


Quality of 
Steam 

After. 


Efficiency. 


B 


Per Cent. 
87.0 
90.1 
89.6 
90.6 
88.4 
88.9 


Per Cent. 

98.8 
98.0 
95.8 
93.7 
90.2 
92.1 


Per Cent. 
90.8 
80.0 
59.6 
33.0 
15.5 
28.8 


A 


D 

C 


E 


F 



Tests with Varying Moisture. 



Quality of 
Steam Before. 



Per Cent. 

66.1-97.5 

51.9-98 

72.2-96.1 

67.1-96.8 

68.6-98.1 

70.4-97.7 



Quality of 
Steam After. 



Per Cent. 

97.8-99 

97.9-99.1 

95.5-98.2 

93.7-98.4 

79 . 3-98 . 5 

84.1-97.9 



Average 
Efficiency. 



Per Cent, 



87. 
76. 
71. 
63. 
36. 



28.4 



300. Classification of Separators. — Separators are based on one or 
more of the following principles of action: 

1. Reverse current. The direction of the flow is abruptly changed, 
usually through 180 degrees. This causes the water in the steam, on 
account of its greater specific gravity, to be thrown into a receiving 
vessel, while the steam passes on in a reverse direction. 

2. Centrifugal force. A rotary motion is imparted to the steam 
whereby entrained water particles are eliminated by centrifugal force. 

3. Baffle plates. The flow is interrupted by corrugated or fluted 
plates to the surfaces of which the water particles adhere and from 
which they fall by gravity to the well below. 

4. Mesh. The separation is brought about by mechanical filtration 
through screens or meshes. 

The following outline shows the classification of typical, separators, 
in accordance with the above principles : 

* See Power, May 11, 1909, p. 834. 



SEPARATORS, TRAPS, DRAINS 



577 



Live-steam separators 



Exhaust-steam separators 



Reverse current 



Centrifugal 



Baffle plate 



Mesh 



Hoppes, Fig. 311. 
|Stratton, Fig. 312. 
Keystone, Fig. 313. 
Mosher. 
Robertson. 
Bundy, Fig. 314. 
Austin, Fig. 315. 
Detroit. 

Direct, Fig. 316. 
Potter mesh. 



( Jacketed baffle Baum, Fig. 317. 

(Absorption Loew, Fig. 318. 



301. Reverse-Current Steam Separators. — Fig. 311 shows a section 
through a Hoppes steam separator and illustrates the principle of 
reverse-current separation. Steam may 
flow through in either direction. Both 
the inlet and outlet ports are surrounded 
by gutters C, C, partly filled with water, 
which intercept the moisture following 
the surface of the pipe, while the down- 
ward plunge of the steam throws the 
entrained water to the bottom of the 
separator. The condensation is carried 
from the troughs by pipe P to the well 
below, from which it is trapped at D in 
the usual way. The velocity of the 
steam in passing through this separator 
is greatly reduced to prevent the steam 
from taking up the water in the bottom 
of the well. This is brought about by 
increasing the area of the passage through 
the separator. 

Fig. 312 gives a sectional view of a Stratton separator, which, though 
primarily of the reverse-current type, embodies also the principle of 
centrifugal force. The separator consists of a vertical cast-iron cylinder 
with an internal central pipe C extending from the top downward 
for about half the height of the apparatus, leaving an annular space 
between the two. The current of steam on entering is deflected by a 
curved partition and thrown tangentially to the annular space at the 
side, near to top of the apparatus. It is thus whirled around with all 
the velocity of influx, producing the centrifugal action which throws 
the particles of water against the outer cylinder. These adhere to the 
surface, so that the water runs down continuously in a thin sheet around 




Fig. 311. 



Hoppes Steam Sepa- 
rator. 



578 



STEAM POWER PLANT ENGINEERING 



the outer shell into the receptacle below. The steam, following in a 
spiral course to the bottom of the internal pipe, abruptly enters it, and 
passes upward and out of the separator without having once crossed 
the stream of separated water. The rapid rotation of the current of 
steam imparts a whirling motion to the separated water which tends to 
interfere with its proper discharge from the apparatus. The separator 
has therefore been provided with wings or ribs E projecting at an acute 
angle to the course of the current, which have the effect of breaking up 
this whirling motion and allowing the water to settle quietly at the 
bottom, whence it passes off through the drain pipe D. 





Fig. 312. Stratton Steam Separator. 



Fig. 313. Keystone Steam Separator. 



303. Centrifugal Steam Separators. — Fig. 313 shows a section 
through a Keystone or Simpson's centrifugal separator. The separator 
consists of a cast-iron cylinder with vertical pipe C extending down- 
ward about two-thirds of the whole length; this pipe has a thread or 
screw wound spirally around it, the space between the threads being 
somewhat greater than the area of the steam pipe. The steam passing 
around the spiral course causes the water to be thrown against the 
outer walls by centrifugal force, while the dry steam passes through 



SEPARATORS, TRAPS, DRAINS 



579 





Fig. 314. Bundy 
Steam Separator. 



the small holes in the central pipe. The water passes down the outer 
walls, where its motion is arrested by obstructing ribs E, and is thence 
carried away by a drip pipe D to a suitable drain. 

303. Baffle-Plate Steam Separators. — Fig. 314 gives an interior 
view of a Bundy separator and illustrates the application of baffle 
plates for live steam separation. This separator 
consists of a rectangular cast-iron casing with a 
cylindrical receiver beneath it. Directly across the 
steam passage are baffle plates corrugated for the 
reception of entrained water. The plates consist 
of vertical castings, each containing a main artery 
or channel which leads directly to the receiver. 
The fronts of the plates are flat, with a series of 
recesses sloping inwards and downwards, terminat- 
ing in an opening of capillary size leading to the 

main artery. The plates 
are staggered, so that 
the steam must impinge 
against all of them in its 
passage. The particles 
of water adhere to the plates, collect, and 
fall by gravity into the receiver. The 
flanges at the bottom constrict the opening 
of the reservoir so as to prevent the steam 
from picking up any portion of the water. 
Fig. 315 shows a section through an 
Austin separator and illustrates another 
class embodying the fluted baffle plate 
principle. The steam in passing through 
the chamber impinges against the fluted 
baffle plate B. The moisture adheres to 
the surfaces, collects, and trickles along the 
corrugations to the bottom of the well. 
These corrugations are formed in such a 
manner that the steam cannot come in 
contact with the water particles after 
they have been once eliminated. A per- 
forated diaphragm D prevents the water in the well from coming 
in contact with the steam. The current of steam is also reversed, 
thus giving additional separating properties to the apparatus. 

304. Mesh Separators. — Fig. 316 shows a section through a " direct " 
separator, illustrating the principle of mesh separation. These separa- 



Fig. 315. 



Austin Steam Sepa- 
rator. 



580 



STEAM POWER PLANT ENGINEERING 



tors are made with steel bodies and cast-iron heads and bases, in all 
sizes up to six inches inclusive, the larger sizes being constructed of 
cast iron or boiler plate. The cone C, perforated lining E, and dia- 
phragm S are made of cold-rolled copper; the cone is a substantial 

gray-iron casting, resting on three cast- 
iron supports hooked over the top of 
inner pipe as indicated. The method 
of operation is as follows: The accumu- 
lated moisture around the walls of the 
steam pipe is caught by the upper edge 
of cone C and carried down back of 
lining E to the water chamber. The 
current of steam entering the separator 
impinges upon the conical surface, which 
is composed of solid plate covered 
with sieve S, through which water may 
freely pass but from which it cannot 
readily escape. Passing through the 
sieve and depositing on the solid surface 
of the cone 0, this water is carried by 
conductors P to the water chamber. 
Perforated lining E permits the moisture 
content of the steam to pass through the 
opening to the water below and prevents 
it from coming in contact again with the 
current of steam. A trough is provided 
at the lower edge of the inverted cup 
which leads all the water that may ad- 
here to it to the water chamber. The 
steam flows through the passages indicated by arrows and is sub- 
jected to a whipsnapping action which tends to throw off any re- 
maining moisture. The perforated plate D prevents the steam from 
picking water out of the water chamber. 

305. Location. — Live-steam separators may be located 

1. Inside the boiler. 

2. Between boiler and engine. 

3. At the steam chest. 

Where the steam pipe is very short, and particularly in marine and 
locomotive work where the tossing of the boiler induces excessive 
priming, the separator may be placed inside the boiler and its 
function becomes that of a dry pipe. In this location it prevents the 
water due to foaming and priming from passing to the engine, and 




Fig. 316. " Direct " Steam Sepa, 
rator. 



SEPARATORS, TRAPS, DRAINS 581 

reduces condensation in the pipe by supplying dry steam. The 
" Potter mesh " and the " De Rycke centrifugal " are types of sepa- 
rators designed for this service. 

The arrangement of separator between engine and boiler, other than 
at the throttle, or inside the boiler is sometimes necessary for economy 
of space. Where possible, however, the separator should be placed 
close to the steam chest. 

Current practice recommends that a receiver separator, which is an 
ordinary separator with a volume of two to four times that of the 
high-pressure cylinder, be placed close to the engine if the load is 
intermittent or sharply fluctuating. This forms a cushion for absorb- 
ing the force of the blows caused by cut-off, delivers steam at a prac- 
tically uniform pressure, and reduces the vibration of the piping to a 
minimum. It also provides a reservoir for sudden demands made by 
the engine. Smaller pipes and higher velocities may be used with this 
arrangement. 

306. Exhaust-Steam Separators and Oil Eliminators. — The function 
of an exhaust-steam separator is the removal of cylinder oil from the 
steam exhausted by engines and pumps. In plants where exhaust 
steam is used for heating it is quite essential to remove the oil from 
the steam before it enters the heating system, for the oil not only 
reduces the efficiency of the radiators by coating them with an excel- 
lent non-conducting film but is an element of danger to the boiler 
itself. In condensing plants the separator will prevent the oil from 
fouling the condenser tubes and those of the vacuum heater if one is 
installed; this is an important factor, since the oil or grease lowers the 
efficiency of the heat transmission. 

In a general sense a live-steam separator is also an oil eliminator, and 
all the separators previously described perform this function to a cer- 
tain extent, since the underlying principles governing the elimination 
of oil from exhaust steam are similar to those employed in removing 
water from steam. Most of the separators described above are also 
designed, in lighter form, as oil eliminators, but by far the greater 
number are based on the fluted baffie plate principle, of which the Hine, 
Bundy, Cochrane, Utility, Peerless, and Keiley are well-known examples. 
This type of oil separator will eliminate a considerable portion of the 
oil in the steam provided the baffle plates or corrugated surfaces are 
frequently cleaned. 

The following is taken from the report of Professor R. Burnham of 
the Armour Institute of Technology on the test of a six-inch horizontal 
oil separator of the baffle-plate type: 

" For purposes of test the separator was placed in the exhaust line 



582 



STEAM POWER PLANT ENGINEERING 



of a 9 x 18 x 24 cross compound Corliss engine running under its maximum 
load at 80 pounds pressure and exhausting into a Wheeler surface 
condenser against 26 inches vacuum. 

" Cylinder oil was fed through the lubricators of the high and low 
pressure cylinders at the rate of from 5 to 20 drops per minute, a record 
being made of the exact quantity of oil fed per hour. The separator 
was so arranged, by means of a receiver connected to the air pump, that 
the accumulation of oil and water could be readily trapped from it at 
any time. In order to determine the quantity of oil given up by the 
condenser, and not properly charged against the separator, each series 
of efficiency tests was preceded by a run of three hours during which 
time no oil whatever was fed to the cylinders. During the last hour a 
record was made of the weight of steam used and a sample of the con- 
denser discharge retained for analysis. 

" The efficiency tests were made by feeding at an excessive rate 
through the lubricators as described above, and when conditions became 
practically constant, records were made for one hour, of the weight of 
oil used, weight of condensed steam, and drain from separator. Samples 
of the two latter were retained for analysis and the percentage of oil in 
them accurately determined, correction being made for the oil given up 
by the condenser. A second series of tests was made exhausting at 
atmospheric pressure. The results obtained are tabulated below. 



Oil in condensed steam with no oil feeding. 

(Charged to condenser.) Pounds per hour 

Oil fed to cylinder, pounds per hour 

Steam condensed per hour, pounds 

Oil caught by separator, per hour, pounds A 
Oil in condensed steam (corrected), pounds 

per hour .B 

Percentage of oil in condensed steam by 

weight, per cent 

Efficiency of separator, percent . „ ... 



Exhausting into 
26-inch Vacuum. 



.051 .057 .0559 

.401 .562 .934 

1000 1120 1096 

.341 .450 .743 

.009 .010 .0096 

.0009 .001 .00088 

97.4 97.8 98.8 



Exhausting at 

Atmospheric 

Pressure. 



.0353 .0340 

.621 .710 

905 872 

.552 .583 



.0071 



.0050 



.00078 .00057 
98.7 99.1 



" There was practically no free oil on the surface of the condenser 
discharge in any case, the small quantity of oil which passed the separa- 
tor (from 5 to 10 parts in a million of water by weight) existing as an 
emulsion, imparting a slight milky color to the water." 

It is a well-established fact that oil can be more effectually removed 
from wet than from dry steam, and some makers, notably the Austin 



SEPARATORS, TRAPS, DRAINS 



583 



FROM ENGINE 



Separator Company, inject a cold-water spray into the separator cham- 
ber. A similar result is brought about in the Baum separator, Fig. 317, 
in which the corrugated baffle plate is hollow and cold water is forced 
through the chamber thus formed. Referring to Fig. 317 : The diverged 
baffle plate forms the wall of a chamber in which cold water is con- 
tinually circulated. This circulation 
causes moisture to appear on the baffle 
plate surface. The particles of oil, 
coming in contact with this moist sur- 
face as the steam current is diverged, 
adhere to it and fall by gravity into the 
well below, where they are completely 
isolated from the purified steam. A large 
portion of the oil and water, however, 
does not enter the separator at all but 
is caught by the inside ledge near the 
junction of the exhaust pipe and the 
separator. The oil and condensation 
which are carried along the bottom of 
the pipe come in contact with this ledge 
and are carried directly to the outlet 
pipe. 

A very successful method of removing 
oil from steam is to project the steam on 
to the surface of a body of water. The 
water may be hot or cold and will hold 
the oil if it once reaches the surface. It 
is essential, however, to reduce the velo- 
city of the steam as it passes on its way 
to the outlet. Baldwin's grease separator is based upon this principle. 
(Baldwin on Heating, p. 234.) 

The most efficient method of removing oil is by combined filtration 
and absorption. {Engineering News, May 22, 1902, p. 406.) A large 
chamber filled with coke, brick, broken tile, or other absorption material 
is placed in series with the exhaust pipe. The steam passing through 
this chamber is entirely freed from oil and moisture, provided the absorb- 
ing material is sufficient in quantity and is replenished as soon as it 
becomes saturated with oil. The annoyance attending the removal 
and replenishing of the absorbing material at frequent intervals and 
the great size of the apparatus are serious drawbacks. An example of 
this system of purification in which many of the objectionable features 
are reduced to a minimum is the Loew grease and oil separator, Fig. 318. 




Fig. 317. 



DRAIN 

Baum Oil Separator. 



584 



STEAM POWER PLANT ENGINEERING 



The exhaust steam enters the chamber at the top, strikes a large 
deflecting plate shaped like an inverted V, and permits part of the 
condensation and oil to be drawn off by the drain pipe. The steam 
then rises and is deflected, as indicated, against a series of shelves 
filled with fibrous material covered with coarse wire screens. The 
grease is removed from each shelf by suitable drains. This apparatus 

is sectional and any number of sections 
may be added without affecting the 
rest. 

In a non-condensing plant where the 
exhaust steam is used for heating pur- 
poses the oil separator is ordinarily 
placed in the main exhaust pipe just 
before it enters the heating system. 
Where several branches enter one main 
it is not customary to place a sepa- 
rator in each branch, one large separator 
located as above being sufficient. 

In condensing plants oil separators 
are seldom installed except where sur- 
face condensers are used, in which 
case the separator may be placed 
anywhere between the engine and 
condenser. In case a vacuum heater 
is used the separator may be placed 
on either side of the heater, depend- 
ing upon the type of separator. If the separator is of the " jacket 
cooling " or " spray " type, it may be placed between the engine and 
the vacuum heater; if. however, it is of the " baffle-plate " type, the oil 
will be more efficiently removed if the separator is placed between 
the heater and condenser so that it will get the benefit of the 
moisture formed in the heater. In the latter location, however, 
the separator will not prevent the oil from fouling the heater 
tubes. 

Where a jet condenser is used and water is taken from the hot well, 
the hot well itself acts as an oil separator. (Trans. A.S.M.E., 24- 
1144.) 

All separators, steam and oil, should be provided with gauge glasses 
and should be thoroughly drained and the drainage should be 
automatic. 

307. Exhaust Heads. — The function of the exhaust head is the 
elimination of oil and water from steam exhaust before permitting it 




Fig. 318. Loew Grease Extractor. 



SEPARATORS, TRAPS, DRAINS 



585 



to be discharged into the atmosphere. Unless removed, the water 
and oil rot the roofs and walls in summer and pollute the atmosphere 
surrounding the plant. The exhaust head also acts as a muffler 
reducing the noise of the escaping steam. Exhaust heads are built 

on the same principle as steam and 
oil separators and most separator 
builders manufacture them. Fig. 
319 shows a section through a typ- 
ical exhaust head. The condensa- 
tion is ordinarily drained to waste, 
though with proper purification it 
may be returned to the boiler. 
With an efficient oil separator in 
the exhaust line the condensation 
in the exhaust head may be returned 
directly to the boiler without further 
purification. 

Live-steam separators are propor- 
tioned so that it is only necessary, in 
the average installation, to specify 
the size of pipe, the type of engine, 
the steam pressure, and the style, 
whether horizontal or vertical. Gauge 
glasses, gauge cocks, and companion 
flanges are usually provided by the 
maker. In some cases the capacity 
of the reservoir is also specified. In 
specifying oil extractors the following additional data are necessary 
for an intelligent choice : the number of engines and pumps exhausting 
into the line, the location of the separator, the steam pressure, velocity, 
and the quality and quantity of cylinder oil used. A guarantee of 
efficiency and of material and workmanship is often demanded. 

i 
REFERENCES. 

Water and Oil Separators: Am. Elecn., Jan., 1900. 

Steam Separators: Am. Elecn., June, 1905. 

" Dry Steam ": Goubert Mfg. Co., 85 Liberty St., N.Y. (Catalogue). 

Cochrane Separator: Harrison Safety Boiler Works, Philadelphia, Pa. 
(Catalogue). 

Bundy Separator: A. A. Griffin Iron Co., N.Y. (Catalogue). 

A Bad Case of Discharge Water with Steam from Water-Tube Boilers: Trans. 
A.S.M.E., Vol. 26. 

Location of Steam Separators: Power, Oct., 1904. 




Fig. 319. A Typical Exhaust-Head. 



586 STEAM POWER PLANT ENGINEERING 

Condensing Exhaust Head: Eng. News, Vol. 49, p. 419; Eng. Rec, Vol. 40, 
p. 177. 

Experiments in Separating Oil from Condensed Steam: Eng. News, May 22, 1903; 
Eng. Rec, April 27, 1901; Engr., Lond., March 12, 1897, Oct. 20, 1905; Power, 
Aug., Sept., 1896, May, 1903; Heating and Ventilation, Feb., 1897; Trans. A.S.M.E. 
Vol, 17, p. 295. 

Oil Separators: Baldwin on Heating, pp. 233-237. 

Oil Separation in a Combination Engine and Turbine Plant: Power, Oct., 1906. 

Test of a Cochrane Steam Separator: Power, April, 1898. 

Test of Lippincott Separator: Engr. U.S., Aug., 1902. 

Test of a Linstrum Steam Separator: Engr. U.S., June 15, 1904. 

Test of an Austin Steam and Oil Separator: Trans. A.S.M.E., Vol. 20, p. 489. 

Test of a Detroit Live Steam Separator: Engr. U.S., April 15, 1904; Power, Jan., 
1902. 

Tests of Direct Separators, Oil and Steam: Direct Separator Co., Syracuse, N.Y. 
(Catalogue). 

Tests of Six Steam Separators: Power, July, 1891, p. 9; Engr. News, Vol. 26, 
p. 233. 

Test of a " Utility " Oil Eliminator: Engr. Rec, May 2, 1903; Engr. Rev., May, 
1903. 

The Hot Well as an Oil Extractor: Trans. A.S.M.E., 24-1144. 

308. Drips. — No matter how thoroughly a steam pipe or reservoir 
may be covered with insulating material considerable condensation 
takes place. With the best covering this loss approximates one-sixth 
of a pound of steam per square foot of pipe surface per hour for steam 
pressures of one hundred pounds, and runs as high as one pound of 
steam for bare pipes. See Table 80 for results of experiments on the 
loss of heat from bare pipes, and Table 81 for data on the efficiency 
of pipe coverings. In addition to this water of condensation, from \ 
to 2 per cent of moisture is carried over by the steam from the boiler. 
This water, unless thoroughly removed, is a constant source of danger 
to the engines and causes water hammer and leaky joints in the piping. 

A joint on a steam pipe may safely withstand a steam pressure of 
100 pounds without leaking and still leak badly under a water pressure 
of half that amount. This is due to the fact that the steam with its 
high temperature causes the pipe to expand, thus insuring a tight 
joint, while the entrained water (which cools as it collects) causes the 
pipe to contract and allows a leak. 

The entrained water and water of condensation are usually spoken of 
as " drips." Drips may be divided into two classes, low pressure and 
high pressure. 

309. Low-Pressure Drips. — Low-pressure drips include the steam 
condensed in exhaust steam feed heaters of the closed type, exhaust 
steam piping, receiver barrels, steam chests, and exhaust heads. As 






SEPARATORS, TRAPS, DRAINS 



587 



these drips are impregnated with oil and are useless for boiler feed 
without purification, they are usually discharged to waste. Most city 




Fig. 320. Closed Heater Installation for Abstracting Heat from Oily Drips. 

ordinances require the drips to be cooled to 100 degrees F. before 
being discharged into the sewer. In this case they must be first dis- 
charged into a tank and permitted to 
cool. This tank must be vented to the 
atmosphere to prevent back pressure. 
Fig. 320 shows an installation in which 
the heat asbtracted from the drips, etc., 
is used to heat the feed water. The 
drips from the throttle valve and steam 
chest in a non-condensing plant are 
ordinarily discharged into the exhaust 
pipe as shown in Fig. 321. In a con- 
densing plant the throttle drips are 
piped to a trap or to the free exhaust 
pipe. The returns from a steam- 
heating system are sometimes classi- 
Simpie Method of Draining fied as low-pressure drips. They are 
Dn P s - invariably returned to the boiler. 

In small plants all the low-pressure drips may be connected to one 
large pipe and this pipe in turn to a single trap, provided there is but 




Fig. 321 



588 STEAM POWER PLANT ENGINEERING 

little difference in pressure in the various drip pipes. In case of different 
pressures separate leads should be run to waste or traps. 

The drips from the receiver and vacuum heater barrels in a con- 
densing plant are oftentimes under less than atmospheric pressure, and 
sometimes the pressure varies from a slight vacuum to 10 or 20 pounds 
gauge, and consequently cannot be disposed of as described above. 
If possible, the heaters and receivers should be placed so as to drain into 
the condenser (see Fig. 334). Should this arrangement prove impracti- 
cable, the barrels may be drained by a trap especially arranged as shown 
in Fig. 335. 

310. Size of Pipe for Low-Pressure Drips. — In the average exhaust- 
steam feed-water heater one pound of steam in condensing gives up 
approximately 1000 heat units. This will heat about 6 pounds of water 
from 60 to 200 degrees F. Hence the area of the drip which carries the 
water of condensation from the closed heater need be but J- that of the 
feed pipe. In no case, however, should a pipe smaller than J inch in diam- 
eter be used. Should the same pipe be used for both exhaust head and 
heater drips, an area of \ area of feed pipe would prove of ample capacity. 
In practice it is customary to use the size of pipe conforming with the 
outlet furnished by the manufacturer of the apparatus, and only when 
several pieces of apparatus are connected to one main are calculations 
made for the size of this main. 

The drip pipe from the throttle valve is ordinarily J inch diameter 
irrespective of the size of steam pipe; this is also true of the steam-chest 
drip. 

311. High-Pressure Drips. — High-pressure drips consist of those 
which are condensed under practically boiler pressure and include the 
steam condensed in steam pipes, cylinder jackets of engines, reheating 
coils of receivers, and separators. Being free from oil and containing 
considerable heat, they are usually returned to the boiler. Drips may 
be returned to the boiler automatically by means of 

1. Steam traps. 

2. Holly steam loop. 

3. Pumps. 

312. Classification of Steam Traps. — Steam traps may be divided 
into two classes, depending on their use, — return and non-return. 
Both of these two classes may be subdivided into five types according 
to the principle of operation, viz. : 

I. Float. III. Bowl. 

II. Bucket. IV. Expansion. 

V. Differential. 






SEPARATORS, TRAPS, DRAINS 



589 



Return Traps. 

Traps which receive the condensed steam and return it to a boiler 
having considerably higher pressure than that acting on the returns 
are known as return trays. They are made in a great variety of styles. 
The general principle of operation is shown in Fig. 331 and described in 
paragraph 318. 

Non-Return Traps. 

Non-return traps, as the name implies, are used where the water of 
condensation is not returned to the boiler but is discharged into any 
receptacle having less than boiler pressure. 



CLASSIFICATION OF A FEW WELL-KNOWN STEAM TRAPS. 

Float 

Bucket 

Dump 



Steam Traps 



Expansion 



Differential 



(McDaniel. 
( Cookson. 

!Acme. 
Albany. 
!Bundy. 
Morehead. 



Metal 



Volatile-Fluid 

(Flinn. 
(Siphon. 



{Columbia. 
Geipel. 
! Dunham. 
Heintz. 




Fig. 322. McDaniel Float Trap. 

313. Float Traps. — Fig. 322 shows a section through a McDaniel 
improved trap, illustrating the principles of the float type. A hollow 
sphere C of seamless copper pivoted at E rises and falls with the change 
of water level in the vessel. The discharge valve M is operated by the 



590 



STEAM POWER PLANT ENGINEERING 



float. When the trap is empty the float is in its lowest position and 
the discharge valve is closed. Water of condensation flows into the 
trap by gravity through opening D to a certain depth, when the float 
opens the discharge valve and the steam pressure acting on the 
surface of the water forces it through outlet S to tank or atmos- 
phere. After the water is discharged the float closes the valve and 
permits the condensation to collect again. A gauge glass indicates the 
height of water in the chamber. 

Unless float traps are well made and proportioned there is a danger 
of considerable steam leakage through the discharge valve, due to 
unequal expansion of valve and seat and the sticking of moving parts. 
The discharge from a float trap is usually continuous, since the height of 
the float, and consequently the area of the outlet, is proportional to the 
amount of water present. When the trap is working lightly, this 
adjustment is apt to throttle the area and create such a high velocity of 
discharge as to cause a rapid wear of valve and seat. This defect is 
more or less evident in all steam traps discharging continuously. For 
this reason all wearing parts should be accessible and readily replaced. 




Fig. 323. Acme Bucket Trap. 

314. Bucket Traps. — Fig. 323 shows a section through an " Improved 
Acme " steam trap. The water of condensation enters the cast-iron 
vessel at A, filling the space D between the bucket E and the walls of 






SEPARATORS, TRAPS, DRAINS 



591 



the trap. This causes the bucket to float and forces valve V against 
its seat (valve V and its stem being fastened to the bucket as indicated) . 
When the water rises above the edges of the bucket it flows into it and 
causes it to sink, thereby withdrawing valve V from its seat. This 
permits the steam pressure acting on the surface of the water in the 
bucket to force the water through the annular space H to discharge 
opening G. When the bucket is emptied it rises and closes valve V 
and another cycle begins. By closing valve R the trap is by-passed 
and the condensation blows directly through passage C to discharge G. 
The discharge from this type of trap is intermittent. 

315. Dump or Bowl Traps. — Fig. 324 shows sections through a 
Bundy bowl trap of the " return " type. The water enters the bowl 




Fig. 324. A Typical Tilting Trap. 



through trunnion D and rises until its weight overbalances counter- 
weight E and the bowl sinks to the bottom. As the bowl sinks, arm 
G, which is a part of the bowl, rises and engages the nuts N on valve 
stem H and opens valve /, thus admitting live steam pressure on to the 
surface of the water. The trap then discharges like all others. After 
the water is discharged weight E sinks and raises bowl A, which in turn 
closes valve /, and the cycle begins again. Bowl traps are necessarily 
intermittent in their discharge. 

Fig. 335 shows the application of a bowl trap to a receiver where the 
drips are under a vacuum, and Fig. 336 a similar application to an 
engine receiver where the pressure varies from less than atmospheric 
pressure to a pressure of 40 or 50 pounds. 



592 STEAM POWER PLANT ENGINEERING 

316. Expansion Traps. — Expansion traps may be divided into two 
groups : 

(1) Those in which the discharge valve is operated by the relative 
expansion of metals and 

(2) Those in which the action of a volatile fluid is utilized. 
Expansion traps will never freeze, as they are open when cold and 

all the water drains out before the freezing temperature is reached. 

Since traps of this type have little capacity for holding water, 5 to 
10 feet of pipe should be provided between the trap and the pipe to be 
drained in order that the condensation may collect and cool. 

Fig. 325 shows the general appearance of a Columbia expansion 
trap in which the valve is operated by the expansion of metallic tubes. 



INLET 




D 

OUTLET 

Fig. 325. A Typical Expansion Trap. 

Water gravitates to the trap through opening marked " inlet/' passes 
through brass pipe 0, then downward to the main body of the valves 
and back to outlet valve C. Below pipe and parallel to it is an 
iron rod S, at the end of which is the support or fulcrum of lever R. 
The lower end of this lever is connected to the stem of the valve C, so 
that any movement of the lever is communicated to it. When the trap 
is cold, valve C is open and all water of condensation passes out. The 
moment steam enters the pipe it expands. The amount of expansion 
is multiplied several times by the action of the lever R, so that the 
movement of the valve is much greater than the expansion of the 
pipe 0. The compensating spring D prevents the brass tube from 
damaging itself by excessive expansion. Lever A permits the trap 
to be blown through by hand. 

Fig. 326 shows a section through a Geipel trap in which the valve 
is operated directly by the expansion of two metallic tubes and the 
movement is not multiplied by levers as with the Columbia. The 
lower or brass pipe constitutes the inlet, and is connected to the vessel 
to be drained; the upper or iron pipe is the outlet for discharge. The 
two pipes form the sides of an isosceles triangle, the base F of which is 



SEPARATORS, TRAPS, DRAINS 



593 



rigid, while the apex A is free to move in a direction at right angles to 
the linear expansion of the tubes. When cold, the brass pipe is con- 
tracted and the apex, in which the valve seat is placed, is moved down 
so that the valve is open and the water is discharged. As soon as steam 




Fig. 326. Geipel Expansion Trap. 

enters the brass pipe the latter expands and forces the valve seat 
against the valve. The trap may be adjusted for any pressure by 
means of the lock nuts E. When it is desired to blow through, the 
valve may be operated by hand by pressing the lever. 

Fig. 327 shows a section through a Dunham trap. It operates upon 
the expansion principle, utilizing a fluid of a volatile character as its 




v~> 



Fig. 327. Dunham Expansion Trap. 



motive force. The corrugated bronze disk B is filled with a volatile 
fluid, and expands and contracts according to the pressure exerted by 
the fluid. The water enters at the top, surrounds disk B and passes 
through valve opening D to discharge outlet at E. As soon as steam 
strikes the disk B the volatile fluid flashes into a vapor and causes the 
disk to expand. This expansion forces valve D against its seat and 
the discharge ceases. The valve will remain closed until the con- 
densation collects and cools the disk B, which then contracts, opens 



594 



STEAM POWER PLANT ENGINEERING 



the valve, and condensation enters as before. The adjustment, how- 
ever, is such that the discharge may be made continuous instead of 
intermittent. 

The Dunham trap is claimed to be the smallest trap of its capacity 
on the market. The 1-inch size, having a capacity for draining 
10,000 lineal feet of 1-inch pipe under 60 pounds pressure, weighs 
but 5 pounds and may be connected to the pipe line as if it were a 
globe valve. 

Fig. 328 shows an internal view of a Heintz steam trap. This 
works on the principle of the volatile fluid expansion trap but in 




Fig. 328. Heintz Expansion Trap. 



a different manner from any of those described above. The requisite 
movement is obtained by the elongation and contraction of the 
extremities of a bent metallic tube T filled with a highly volatile 
fluid. This tube is inclosed in a cast-iron box and presses against 
the point of regulating screw P. The other extremity of the tube 
carries the valve and is free to move under the action of the 
variations of temperature. Spring S has no connection with the 
action of the trap. It is used as a simple means of holding one 
end of the expansion tube on its pivot. The trap operates as follows: 
Water enters at /, surrounds the tube T and passes through the 
valve to the discharge outlet O. As soon as steam enters the chamber 
the volatile fluid in the tube flashes into a vapor and the pressure 
thus created tends to straighten out the tube; this forces the valve 
against its seat and the discharge ceases. As the trap cools the tube 
returns to its normal position and the discharge valve is opened, thus 
permitting the condensation to drain out. The adjustment permits of 
continuous or intermittent discharge and of variable pressures. 

317. Differential Traps. — Fig. 329 shows a cross section through 
a Flinn differential trap. The column of water X acting on dia- 
phragm D closes valve V. The water entering pipe E and the 
action of the spring equalize column X and open the valve. 



SEPARATORS, TRAPS, DRAINS 



595 



Describing the action in further detail, the water of condensation 
enters at A, fills lower chamber Y, pipe X, and receiving chamber C 
up to the level of the top of pipe E. This column of water acting on 
the under side of the diaphragm D forces the valve to its seat against 
the counter pressure of the spring S. Any additional water that 
enters the trap overflows through pipe E, filling chamber F and pipe E 
to a point about midway of its height, where the effect of the 
column of water in pipe X is balanced. The pressure on each side of 
the diaphragm is then equal, the short column in pipe E, aided by 
the spring, balancing the pressure of the longer column in pipe X. 

Any further increase in the height 
of the water in pipe E causes a 
depression of the valve V, which 
allows water to escape until the 
column has fallen to a level a 
little below the middle of pipe E, 
when this valve closes again. This 
action is repeated at intervals 
according to the quantity of water 
entering the trap. 
So long as the water 
keeps coming in 
sufficiently large 
quantities the valve 
remains wide open. 
Fig. 330 gives a 
general view of a 
siphon trap which 
is much used in 
draining low-pres- 
sure systems, as, 
for example, the 
separator in an ex- 
haust steam heat- 
ing system. It con- 
sists essentially of 
two legs A and B, which may be close together or any distance apart 
but the lengths of which must be sufficiently great to prevent pressure 
acting through pipe / from forcing the water out of B. C is a vent pipe 
extending to the air to prevent siphoning; is the discharge for the 
condensed steam. In ordinary operation the leg B is filled with water 
which is constantly overflowing, and A with steam and water, the 




FP° 



Y 

Fig. 329. 



V Y 

Flinn Differential Trap. 



G 

Fig. 330. Simple 
Siphon Trap. 



596 



STEAM POWER PLANT ENGINEERING 



EQUALIZING VALVE 




total pressure in both legs being equal. The siphon trap is applicable 
for low pressure only, as it requires approximately 2.3 feet of vertical 
space E for each pound per square inch pressure in the pipe. The 
maximum allowable head is represented by vertical distance N. 

318. Location of Traps. — Wherever possible a trap should be 
located so that the condensation will flow into it by gravity. This 
will insure positive drainage. Sometimes, however, the coils, cylin- 
ders, or pipes to be drained are located in a pit or trench or lie on a 
basement floor where it is impossible to set the trap so as to receive 
the drains by gravity without placing it in an inaccessible position. 
With very low pressures this is often unavoidable, but with pressures of 
five pounds or more the trap may be placed above the point to be 
drained. If a trap is set in an exposed place a drain should be pro- 
vided at the lowest point to free the pipe of water when steam is shut 
off. A dirt catcher or strainer should be placed in the pipe leading to 
the trap to prevent scale, etc., from reaching the valve. All pockets 
and dead ends should be drained, and no condensation should be 
allowed to accumulate. High and low-pressure drips should be kept 
separate. All tanks should have gauge glasses. 

Fig. 331 shows the ap- 
f \ STEA " SUPPLY — r ^ plication of a float trap 

for automatically returning 
water to the boiler. For 
this purpose the trap must 
be placed three feet or 
more above the water line 
in the boiler, so that the 
water may gravitate to it. 
Water is forced into the 
trap from the returns 
through pipe A until it 
reaches a level where the 
float opens the equalizing 
valve V and permits steam 
from the boiler to enter the 
trap, thus equalizing the 
pressures. The water then 
flows into the boiler by 
gravity through check valve D. At the end of discharge the float 
closes the equalizing valve and another cycle begins. Check valve C 
prevents the water from being forced back to the return pipe. If the 
pressure in the return pipe A is not sufficient to force the water into 



F£> 



RETURN TRAP 



■=-=--WATER LEVEL 



Fig .331. Return Trap . 






SEPARATORS, TRAPS, DRAINS 



597 



the trap, a pump or another trap may be used to effect this result. 
Practically any high-pressure trap may be converted into a return trap 
by proper installation and an " equalizing " valve. 




TRAP TRAP 



Fig. 332. Drainage System for Jackets and Receivers of Triple Expansion Pumping 

Engines. 

Figs. 332 and 333 show different applications of steam traps to the 
receiver coils and jackets of triple expansion pumping engines. The 
drawings are self-explicit. 



SEPARATOR 



tn 




FEED TANK 



Fig. 333. Drainage System for Jackets and Receivers of Triple Expansion Pumping 

Engines. 

319. Drips under Vacuum. — Conditions frequently make it neces- 
sary to remove condensation from apparatus working under a vacuum, 
as, for example, a primary heater. 

The simplest method is to pipe the drips to the condenser and per- 
mit the condensation to gravitate to it as in Fig. 334. Where this is 
impracticable, as in an installation with the condenser above the 



598 



STEAM POWER PLANT ENGINEERING 



heater, a steam trap is usually employed. Fig. 335 shows the applica- 
tion of a Bundy trap to a vacuum or primary heater. A close-fitting 

weighted check valve W, set to 
open outwards prevents intake 
of air through the discharge 
pipe while the trap is filling. 
Connection E is made from 
the vent underneath the valve 
stem V back to the heater so 
as to equalize the pressures. 
The operation is as follows: 
Condensation gravitates from 
the heater through check C to 
the body of the trap, the check 

Fig. 334. Gravity Drainage ; Vacuum Heater. W being closed. When the 

bowl is full enough to overcome 
the weight of the counterbalance, it sinks and opens up the live-steam 
valve V. This admits steam to the trap through pipe D, which in 





Fig. 335. Method of Draining Heater under Vacuum. 

turn closes check C and forces the water past the weighted . check W 
to the discharge tank. After the water is discharged the bowl returns 






SEPARATORS, TRAPS, DRAINS 



599 



to its original position and closes valve V, the weight closes check W, 
the vent check equalizes the pressure in the bowl and heater, and 
condensation gravitates to the trap again. 

320. Drips under Alternate Pressure and Vacuum. — Occasionally 
the load on an engine is of such a character that the pressure in the 
receiver alternates from a pressure of 30 or 40 pounds absolute to a 
vacuum of varying degree. Where the periods of vacuum operation 
are very few and of short duration, as in the average installation, no 
attention is paid to the vacuum and the condensation is removed by 
a trap in the ordinary way. If, however, the periods are of sufficient 
duration and frequency, the ordinary method is not applicable and the 
arrangement shown in Fig. 336 may be used. The trap is placed 




Fig. 336. Method of Draining Receivers under Alternate Vacuum and Pressure. 



below the receiver as indicated. The delivery pipe is provided with a 
weighted check or resistance valve W set so as to open outwards from 
the trap, also a spring water relief valve R. Another weighted check P 
is placed in the line leading from the vent to the atmosphere, and a 
plain check C in the line leading back into the receiver. This arrange- 
ment of valves permits the venting of the trap after discharge and 
effectually excludes air from the trap when there is less than atmos- 
pheric pressure on the receiver. With the relief valve set to open at a 



600 



STEAM POWER PLANT ENGINEERING 



pressure in excess of the maximum receiver pressure it acts as a 
" stop " in the pipe and the water must enter the trap. When the 
trap discharges, the live steam supplied through the pipe attached to 
the steam valve forces the water through the weighted check and relief 
valves into the sewer or receiving tank. When working with a vacuum, 
the pressures in receiver and trap are equalized through the vent con- 
nection and the condensation flows into the trap by gravity. The 
operation of discharge is the same as in the case of pressure. 

321. The Steam Loop. — Fig. 337 illustrates the principles of the 
" steam loop " for automatically returning high-pressure drips to the 




\ HORIZONTAL 



"7- 



cS 



kZ 



o5 



^. 



===L—. ; WATER LEVEL: 



Fig. 337. General Arrangement of the Simple " Steam Loop.*' 

boiler. In the figure the loop is returning the condensation from a 
steam separator to a boiler above the level of the separator. The 
apparatus is very simple, consisting of a horizontal and two vertical 
lengths of plain pipe placed as indicated. Pipes R and B may be cov- 
ered but " horizontal " A is left uncovered, as its function is that of a 
condenser. The operation is as follows: Circulation is first started by 
opening stop valve at the bottom of the drop leg until steam 
escapes. The valve is then closed and the steam in the horizontal A 
condenses and gravitates to the drop leg B. On account of the slight 
reduction in pressure in the horizontal a mixture of spray and steam 
flows from the separator chamber to the horizontal, and, condensing, 
gravitates to the drop leg. The column of water in the drop leg rises 
until its static head balances the difference in pressure in the riser R 
and the horizontal. In other words, a decrease in pressure in the 
horizontal produces similar effects on the contents of the riser and 
drop leg but in a degree inversely proportional to their densities. 



SEPARATORS, TRAPS, DRAINS 



601 




602 



STEAM POWER PLANT ENGINEERING 




Any further accumulation causes an equal amount to pass from the 
bottom of the column to the boiler, since the pressure in the boiler is 
then less than that at the bottom of the column, that is, the steam 
pressure on the top of the water column plus the hydrostatic head H 
is greater than the pressure in the boiler. Once started the process is 
continuous and requires no further attention. 

322. The Holly Loop. — In the application of the steam loop where 
many points requiring drainage are connected to many boilers and 
conditions are more complex, some method other than the simple one 
of radiation may be advisable to secure the necessary lower pressure at 
the top of the loop. Such a method is illustrated in Fig. 338. This 
arrangement differs from the simple loop in that all condensation first 
gravitates to a " Holly " receiver (shown in detail in Fig. 339) before 

passing into the " riser." 
The receiver is placed be- 
low the lowest point to be 
drained and serves as a 
storage for large or unusual 
quantities of water and en- 
ables the riser to act at a 
constant rate independent 

Fig. 339. Holly Receiver. Q f var i a ble discharge into 

the receiver. Furthermore, the lower pressure in the discharge cham- 
ber necessary to secure the lifting of the mingled steam and water 
through the riser, instead of being created by condensation as in the 
simple loop, is produced by a reducing valve B discharging into the 
feed-water heater. The operation of the Holly loop is as follows: 
Circulation is started by opening valve D until steam appears. Valve D 
is then closed and the reducing valve is put into commission. Condensa- 
tion from separators, traps, and pipes gravitates to the " receiver," from 
which it is forced into the " riser " in the form of a spray. The spray- 
ing effect is produced by a series of holes drilled in pipe A, Fig. 339. 
From this receiver the spray and moisture rise to the " discharge 
chamber/' on account of the lower pressure at that point, where 
the steam and entrained water are separated, the water gravitat- 
ing to the bottom of the chamber and thence to the drop leg, and the 
steam discharging through the reducing valve into the heater. The 
principles of operation are exactly the same as in the simple steam 
loop. 

323. Returns Tank and Pump. — Low-pressure drips in connection 
with heating systems may be returned to the boiler along with the 
condensation from the heating system by a combined pump and receiver 



SEPARATORS, TRAPS, DRAINS 



603 



as shown in Fig. 340. The height of water in the tank controls the 
operation of the pump through the medium of a float and throttle valve. 
This combination of float and balanced throttle valve is sometimes 
called a " pump governor." In the illustration the pump forces the 
returns through a closed heater before delivering them to the boiler, 




Drip Trapped 
•to Sewer ^^rr 



Fig. 340. Returns Tank and Pump. 

though they are oftentimes returned directly. The tank is vented to 
the atmosphere to prevent it from becoming " air bound." The cold- 
water supply or make-up water is sometimes discharged into the receiv- 
ing tank as indicated. With open heaters the cold supply is ordinarily 
controlled by a float within the heater itself. 

324. Office Building Drains. — In the power plants of tall office build- 
ings the public sewers are often above the basement level, and it is 
necessary to remove all liquid wastes mechanically. 

The Shone pneumatic ejector has been found to serve this purpose 
effectually. This apparatus is placed in a pit in the basement floor 
into which all sewage, drips from engines, washings from boilers, and 
ground water gravitate, and are automatically discharged into the 
street sewer by means of compressed air. 



604 



STEAM POWER PLANT ENGINEERING 



Fig. 341 gives a sectional view of a Shone ejector of ordinary con- 
struction. It consists essentially of a closed vessel furnished with 
inlet and discharge connections fitted with check valves, A and B, 
opening in opposite directions with regard to the ejector. Two cast- 
iron bells, C and D, are linked to each other, in reverse positions, the 



*•« EXHAUST PI fl A,R SUPPLV 




Fig. 341. Shone Ejector. 

rising and falling of which control the supply of compressed air through 
the agency of automatic valve E. 

The bells are shown in their lowest position, the supply of compressed 
air is cut off from the ejector, and the inside of the vessel is open to the 
atmosphere. The sewage gravitating into the ejector raises the bell C, 
which in turn actuates the automatic valve E, thereby closing the con- 
nection between the inside of the ejector and the atmosphere and open- 
ing the connection with the compressed air. The air pressure expels 
the contents through the bell-mouthed opening at the bottom and the 
discharge valve B, into the main sewer. Discharge continues until the 
level falls to such a point that the weight of the sewage retained in 
the bell Dis sufficient to pull it down, thereby reversing the automatic 
valve. This cuts off the supply of compressed air and reduces the 
pressure to that of the atmosphere. 



SEPARATORS, TRAPS, DRAINS 605 

The positions of the bells are so adjusted that compressed air is not 
admitted until the ejector is full, and is not allowed to exhaust until 
emptied down to the discharge level; thus the ejector discharges a 
fixed quantity each time it operates. 

Two ejectors, each of a capacity suitable for handling the average 
flow of tributary sewage and so arranged that they can work either 
independently or together, are usually installed at each ejector station. 

The main sanitary sewer of the building usually discharges directly 
into the ejectors, the surface water, drips, etc., being collected in a 
neighboring sump. The latter is connected to the sanitary sewer 
through a trap or back-water valve. 



CHAPTER XV. 

PIPING AND PIPE FITTINGS. 

325. General. — The advent of high pressures and superheat is 
responsible for the elimination of many of the older systems of piping, 
the tendency being towards greater uniformity in design, particularly 
in electric central station work. In isolated stations the conditions 
of operation and installation are so variable that each case presents an 
entirely different problem. In any system of piping the fundamental 
object is to conduct the fluid in the safest and most economical manner. 

The material should be the best obtainable and the system so flexi- 
ble that a break-down in one element will not necessitate the closing 
down of the entire plant. On the other hand, flexibility increases the 
number of parts and, unless first cost is of little importance, tends to 
weaken the system as a whole. It is a safe general proposition to say 
that the best pipe and fittings, irrespective of first cost, will prove the 
most economical in the end, but few owners of power plants are will- 
ing to take this view. 

326. Drawings. — An assembly drawing of the entire installation 
giving the location of all valves and fittings is necessary in order to 
avoid interference, and particularly where a number of fittings are to 
be close together. Detailed drawings should also be provided of each 
division of the piping to facilitate installation, as, for example, the 
high-pressure steam, the exhaust steam, the feed water, the condensing 
water, the oil, the heating and the sanitary piping. As a rule, lower 
and more uniform bids will be obtained from an isometric or perspec- 
tive sketch, as in Fig. 254, than from conventional plan and elevation 
drawings, due, no doubt, to the greater ease with which the drawing is 
interpreted. A complete set of specifications for a piping system is given 
in paragraph (415), and illustrates the usual practice along this line. 

327. Materials for Pipes and Fittings. — The following materials are 
used in the construction of pipes for steam, water, and gases. 

Average Bursting Tension. 

Low-carbon or mild steel 65,000 lbs. per sq. in. 

Wrought iron 50,000 lbs. per sq. in. 

Cast iron, high grade 20,000 lbs. per sq. in. 

Cast steel 50,000 lbs. per sq. in. 

Wrought copper 33,000 lbs. per sq. in. 

Brass 18,000 lbs. per sq. in. 

Special alloys and compounds 15,000-60.000 lbs. per sq. in. 

606 



PIPING AND PIPE FITTINGS 607 

Mild Steel. — The greater portion of the piping in the average steam 
power plant is of mild steel, lap or butt welded for high pressures and 
riveted for very low pressures and large diameters. Steel pipe is con- 
siderably cheaper than that manufactured from other material, and 
fulfills practically all requirements for general service. 

Wrought Iron. — Pipes manufactured from puddled iron, though not 
as strong in ultimate bursting strength as low-carbon steel pipes, are 
superior in many ways. Threads are cut more readily and with less 
power. It is more easily bent without injury and offers more resistance 
to corrosion. The life of wrought-iron pipe is greater than that of steel 
pipe under conditions of extreme exposure, or when buried under- 
ground, and it is recommended in all cases where corrosion is apt to be 
severe, as in blow-off pipes, drips, and drains or in pipes not in continuous 
use. Steam pipes well covered and protected from external moisture 
are ordinarily of mild steel, as the conditions do not warrant the use of 
wrought iron, which costs approximately 10 per cent more. Since the 
term " wrought-iron pipe " is used rather loosely in practice, the manu- 
facturer will generally furnish mild steel unless special stress is laid upon 
the term " wrought iron." 

Wrought Iron vs. Steel Pipe: Ir. Age, March 2, 1905, Jan. 18, 1906; Pro. Heat and 
Vent. Engr., Jan., 1900; Mach., Dec, 1903, p. 191; Am. Mfg. and Ir. World, April 
29, 1898; Eng. News, 50-292, 296, 487, 51-62; Eng. Rec, 44-54; Locomotive, Jan. 
1, 1906. 

Corrosion of Pipes: Trans. A.S.M.E., 18-282; Eng. Rec, 42-194, 43-354, 45-584. 

Cast-iron Pipes. — Cast iron is little used for high-pressure steam, 
except occasionally in the construction of headers where a number of 
branches lead into a single pipe, in which case the number of joints is 
greatly reduced and the cost considerably less than for wrought-iron or 
steel pipe with numerous fittings and joints. The chief objections to 
cast iron for high-pressure steam are its weight and lack of homogeneity. 
It is most used in connection with water service and sanitation. 

Cast-iron Pipe: Jour. New Eng. Waterworks Assn., March, 1907; Engr., Lond., 
Nov. 7, 1902, p. 454; Power, Jan., 1904, p. 334; Eng. News, 46-216, 48-193, 51-544. 

Cast-Steel Pipe. — Cast-steel headers are sometimes used in power 
plants for highly superheated steam, since the material is not affected 
by high temperatures to the same extent as cast iron. High first cost 
and the difficulty of securing castings free from blow-holes have pre- 
vented its more general use. 

Copper Pipes. — Copper steam pipes were in common use for many 
years in marine service on account of their flexibility. To increase the 
bursting strength, pipes above 6 inches in diameter were generally wound 



608 STEAM POWER PLANT ENGINEERING 

with a close spiral of copper or composition wire. In recent years 
wrought-iron and steel pipe bends have practically superseded copper 
for flexible connections. As a rule the use of copper pipes should be 
avoided on account of the rapid deterioration of the metal under high 
temperatures and stress variations. The cost is prohibitive for most 
purposes and this alone prevents it from being seriously considered in 
the manufacture of pipe. 

Copper Pipes: Engr., Lond., April 15, 1898, p. 360, Aug. 11, 1893; Engng., April, 
1898; Eng. Rec, July 30, 1898. 

Brass Pipes. — Brass is little used in the construction of pipes on 
account of its high cost. It withstands corrosive action much better 
than iron or steel and is often used in connecting the feed main with 
the boiler drum. Special alloys, nickel steel, " ferrosteel," malleable 
iron, and the like have been used in the manufacture of pipes, and 
possess points of superiority over wrought iron and steel for some 
purposes, as for highly superheated steam, but the cost is prohibitive 
for average steam power plant practice. 

Materials for Fittings. — Elbows, tees, flanges, and similar fittings 
are usually made of cast iron, malleable iron, or pressed steel, though 
cast steel, " ferrosteel," and other steel compounds are used to a limited 
extent. Standard cast-iron fittings are recommended for ordinary 
pressures of 100 pounds per square inch or less, and extra heavy cast- 
iron fittings for higher pressures. Malleable iron fittings are lighter and 
neater than cast iron and are extensively used for small sizes of steam 
and gas pipe. 

Manufacture of Pipe: Sci. Am., Dec. 12, 1903; Mach., Feb., 1904, Dec, 1903, 
p. 191; Eng. News, 50-232, 296. 

328. Size and Strength of Commercial Pipe. — Wrought-iron and 
mild-steel pipe are marketed in standard sizes. Those most commonly 
used in steam power plants are designated as 

(1) Merchant or standard pipe. 

(2) Full weight pipe. 

(3) Large O. D. pipe. 

(4) Extra heavy. 

(5) Double extra heavy. 

Table 76 gives the dimensions of standard " full-weight " pipe, 
which is specified by the nominal inside diameter up to and including 
12 inches and based on the Briggs standard. Pipes larger than 12 
inches are designated by the actual outside diameter (O. D.), and are 
made in various weights as determined by the thickness of metal speci- 






PIPING AND PIPE FITTINGS 609 

fied. Manufacturers specify that " full weight " pipe may have a 
variation of 5 per cent above or 5 per cent below the nominal or table 
weights, but merchant pipe, which is the standard pipe of commerce, 
such as manufacturers and jobbers usually carry in stock, is almost 
invariably under the nominal weight. It varies somewhat among the 
different mills, but usually lies between 5 and 10 per cent under the table 
weight. The smaller sizes of merchant pipe, J inch to 3 inches, are butt- 
welded and the larger sizes are lap-welded. 

Extra heavy and double extra heavy pipe have the same external 
diameter as the standard, but are of greater thickness and hence the 
internal diameter is smaller. Taking the thickness of the standard pipe 
as 1, that of the extra heavy is approximately 1.4 and of the double 
extra heavy 2.8. 

Wrought-iron and steel pipes are ordinarily designed with factors of 
safety of from 6 to 15, with an average not far from 10. The standard 
hydrostatic tests to which the various pipes are subjected at the mills 
are as follows: 

Hydrostatic Pressure, 
Lbs. per Sq. In. 

Standard, butt-welded, |-3 in 600 to 1,000 

Standard, lap-welded, 3-12 in 500 to 1,000 

Extra heavy, butt-welded, f-3 in 600 to 1,500 

Extra heavy, lap-welded, 1^—12 in 600 to 1,500 

Double extra heavy, butt-welded, J- 2 J in 600 to 1,500 

Double extra heavy, lap-welded, 1^-8 in 1,200 to 1,500 

The pressure necessary to burst piping is far above anything likely 
to occur in ordinary practice on account of the thickness of material 
necessary to permit of threading, thus: 

Actual Bursting 

Pressure, 
Lbs. per Sq. In. 

2-inch " standard " mild-steel pipe* 5,800 

2-inch " wrought-iron pipe 4,106 

10-inch " mild-steel pipef 3,000 

10-inch " wrought-iron pipe t 1,900 

10-inch " extra heavy " wrought-iron pipe t 2,700 

* Machinery, December, 1903, p. 192. f Crane Company, Published Tests. 

Riveted Pipes. — .For low pressures and large diameters, pipes are 
constructed of thin sheets of boiler steel with riveted joints, the seams 
being either longitudinal and circumferential, or spiral. Such pipes 
are not necessarily limited to large sizes and low pressures, though this 
is the usual practice. 

Pipe fittings are classed as screwed or flanged. 



610 



STEAM POWER PLANT ENGINEERING 



329. Screwed Fittings, Pipe Threads. — For screw connections the 
ends of pipes and fittings are threaded to conform to the Briggs or 
United States standard system, as shown in Fig. 342. The end of the 




Fig. 342. Standard U. S. Pipe Thread. 

pipe is tapered 1 to 32 with the axis, the angle of the thread being 
60 degrees and slightly rounded at top and bottom. The proper length 
of perfect threads is given by the formula 



T = 



(0.8 D + 4.8) 



(146) 



in which 



T = length in inches. 

D = actual external diameter of the tube, inches. 

n = number of threads per inch. 

The imperfect portion of the thread is simply incidental to the pro- 
cess of cutting. The object of the taper is to facilitate " taking hold " 
in making up the joint. Table 76 gives the number of threads per 
inch for various sizes of standard pipe. When properly constructed a 
screwed joint will hold against any pressure consistent with the 
strength of the pipe. For example, the ultimate bursting strength of a 
" standard" 2-inch pipe is about 6000 pounds per square inch, while 
the stripping strength of the joint (with perfect threads) is 225,000 
pounds. The threads, however, are often poorly cut and the parts 
screwed together improperly cleaned and lubricated, thus causing 
leakage between the threads. 

330. Flanged Fittings. — In cast-iron pipes, valves, tees, and other 
fittings the flange is always a part of the casting, but for joining 
the two ends of a steel or wrought-iron pipe the flanges may be 
fastened to the pipe in a number of ways. Fig. 343, A to H, illus- 
trates methods most commonly used. In A to C the pipes are 
screwed into cast-iron or forged-steel flanges and the two faces, with 
metallic or composition gasket between, are drawn together by bolts. 
A illustrates the most common and inexpensive of flanged joints, 
which requires no special tools and can be made up at the place of 



PIPING AND PIPE FITTINGS 



611 



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•>*rJllOiO00O3C0©ailOC0^'*©'*iO©NNNir3 
OOHiOOSW<NMOTtlN(OH<OHOOOOOO'*iON 


HHHiNn^iocNosHeq^ioosiNioooH^N 


1 


01 

1 


(NtO'-IO)0>Hl005'HlMCOCONOONCOlOtOOO(Ni*iO 
NOSINMOSMr-lffitOCOaitOMONr-ilOOSMNFHlO 
NOHtOINHNOSTliOanOHNTtlOOOsOINNOBO 


Hi-lfqNM'HiO>ONOJO(N'*iONOWNOM©0 


•ssaroptqi, 
Ibuioio^i 


I 

d 


OOOOiHOSW* lfl-*-*NtON(0 0) .-< (N "* «D 
©000SOH«'*^>0O>H(NC0^i000O(N'*© 
OOOHHrtHHHMINNNINMNCOMmP: 










o> 
"S 

s 

[ 


Approxi- 
mate 
Internal 
Diameter. 


■8 

d 

i— i 


T^^tleO^OO THNOONQOCO»lOlOM(NNai 
NCOfflNN^OOrHtOcOcOtiMO^CINOOCOrH 
NM^COOOMOO^OiOOiOOOOOSOSO 




HHHNeq««^^iiotoNNxoH 


C<1 


Actual 
External 
Diameter . 


to 
<D 

X! 
O 

d 


O "^ l>> TtH lO i— i CO t^b- ONMNNiflifliC 
"*lO©QOOCO!OGMi)Om >0 lO©«©«ONNN 




HHrtrHCqiNCCOt^lflKJONOOOSOHCq 




'imnaiui 
imnoio^ 


© 

d 


rHlOO H-* «*» Hn *« rtH»H|N W|(M ~<|<M win 

HHHNNCOMDliJiiOfflNMOSOH 


CN 





612 



STEAM POWER PLANT ENGINEERING 



erection. It gives satisfactory results for pressures of 100 pounds or 
less, but for higher pressures leakage is apt to take place between the 
threads. The flanges are sometimes made with a long thread and a 
recess which can be calked with soft metal. A similar joint is made 



m4* W?i 

SMOOTH FACE 




<^=EP 




U_^j™ pj 0t| 



SHORT HUB 



SCREWED 



S^ 



TONGUED FACE 




SCREWED & PEENED 



RAISED FACE 

T7 1 





^|A1 LONG HUB ls\ 




SHRUNK 



RIVETED 



Fig. 343. Types of Pipe Flanges. 

with the pipe screwed beyond the face of the flange and the two faced 
off together, either plane or as shown in B, which is known as a male 
and female or hydraulic joint. This method forms a very reliable joint, 
since the ends of the pipe bear on the gasket, and the gasket is pre- 



PIPING AND PIPE FITTINGS 613 

vented from being blown out. An objection lies in the difficulty of open- 
ing the line to remove the gasket or replace a fitting. C is a modifica- 
tion known as the tongued and grooved joint, which uses an extremely 
narrow gasket. Such flanges may be subjected to severe strains when 
the bolts are drawn up, owing to the small area of contact. Corru- 
gated copper or steel gaskets are recommended, since soft material is 
apt to be squeezed out. In C the ends of the pipe are peened, which is 
an improvement over the simple screwed joint. D illustrates a shrunk 
joint. The flanges are bored for a shrink fit and forced over the pipe 
when at a red heat. After cooling, the end is beaded over into a recess 
on the face of the flange and a light cut taken from both. H shows a 
modification in which the hub is riveted to the pipe. E illustrates a 
joint constructed by rolling the pipe into a corrugation in the flange. 
The end of the pipe is then faced off flush. 

One of the best commercial joints is illustrated by F and is known 
as the lap joint. The pipe is expanded as indicated and a light cut is 
then taken from the flared ends to insure a tight joint. The flanges 
are loose and permit of considerable flexibility in shifting them through 
various angles. This is sometimes called the Van Stone joint. 

Pipes with flanges welded on the end as in G have proved the most 
reliable of all and though costly are considered the standard for high- 
pressure and high-temperature work. The faces are ordinarily raised 
A to tV m ch inside the bolt holes and ground to a steam-tight fit, so 
that thick gaskets are unnecessary. 

For moderately high pressures and temperatures any of the joints 
when well made will prove satisfactory. For extremely high pres- 
sures and temperatures the lap or welded joints are preferable. 

The comparative costs of various flanges are given in Table 77. 

On July 18, 1894, committees of the American Society of Mechan- 
ical Engineers, of the National Association of Steam and Hot Water 
Fitters, and of Manufacturers met and adopted a schedule for dimen- 
sions of flanges, known as the " A.S.M.E. Master Steam Fitters' " 
flange schedule. Flanges dimensioned in accordance with this schedule, 
Table 78, are applicable for pressures up to 100 pounds per square 
inch. For higher pressures a schedule shown in Table 79 was adopted 
June 28, 1901. 

Pipe Flanges and Fittings: Power & Engr., Mar. 2, 1909, p. 402, June, 1904, 
p. 354, Jan., 1904, p. 52, Oct., 1902, p. 36; Am. Mach., April 18, 1901; Jour. Assn. 
Engr. Soc, Sept., 1904; Jour. Am. Soc. Nav. Engrs., May, 1905; Trans. A.S.M.E., 
20-737, 429, 8-29, 347, 14-49, 7-311; Eng. News, 41-183, 323; Eng. Rec., 41-113, 
44-209. 

Methods of Welding Nozzles, Flanges, etc., to Steam Pipes: Power & Engr., Sept. 28, 
1909, p. 518. 



614 



STEAM POWER PLANT ENGINEERING 



TABLE 77. 
COMPARATIVE COST OF VARIOUS PIPE FLANGE FITTINGS, 12-INCH PIPE. 

(Circular from the Crane Company.) 









.J ^ 

!« 

^ be 
«8 o 

h3 iJ 


"S 3 

|w 

O. O 
1-3 CQ 


C 3 
h3 fc. 


1 


t5 

1 


33 


Cast iron 


% 7.40 

8.70 

9.90 

22.40 

26.40 


$16.00 
18.40 


$18.00 
20.00 








$13.00 
16.00 
18.00 
25.00 
30.00 


$21.00 


Ferrosteel 








23.40 


Malleable iron 


$22.00 








Cast steel 


28.40 
32.40 


34.00 
38.00 


$33.00 
37.00 


$41.00 


33 40 


Weldless steel 


37 40 







Any of the above screwed, shrunk, welded, rolled, or single-riveted flanges can be 
furnished with male or female face at $1 .25 extra. 

The screwed or welded flanges can be furnished with tongued or grooved face at 
$1.25 extra. 

Any of the above screwed, shrunk, or single-riveted flanges can be furnished with 
calking recess at $1.25 extra. 



TABLE 78. 

SCHEDULE OF STANDARD FLANGES ADOPTED JULY 18, 1894, BY A COM- 
MITTEE OF THE MASTER STEAM AND HOT WATER FITTERS' ASSOCIATION, 
AMERICAN SOCIETY OF MECHANICAL ENGINEERS, AND VALVE AND 
FITTING MANUFACTURERS. SUITABLE FOR PRESSURES UNDER 100 
POUNDS PER SQUARE INCH. 









Size of Bolts, 


Size of Bolts, 


Flange 




Width 


Size of Flange, 

Pipe 

SizeX Diam. 


Diameter 
of Bolt 
Circle. 


Num- 
ber of 
Bolts. 


Pressure 
under 80 
Pounds. 


Pressure 80 

Pounds and 

Over. 


Thickness 
at Hub for 
Iron Pipe. 


Flange 
Thickness 
at Edge. 


of 

Flange 
Face. 


2X6 


4f 


4 


4X2 


|X'2 


1 in. 


5 

8 


2 


24X 7 


5* 


4 


IX 21 


1X24 


14 


44 


24 


3 X 7* 


6 


4 


4X24 


1X24 


H 


4 


24 


34X 84 


7 


4 


4X24 


1X24 


14 


if 


24 


4X9 


74 


4 


IX 2| 


IX 2f 


if 


41 


24 


4JX 9i 


71 


8 


IX 3 


|X3 


U 


41 


2| 


5 X10 


84 


8 


1X3 


|X3 


14 


41 


H 


6 XII 


94 


8 


1X3 


|X3 


H 


l 


24 


7 X12i 


10f 


8 


^X oj 


IX 34 


H 


l* 


2| 


8 X13J 


HI 


8 


1X34 


1X34 


14 . 


14 


24 


9 X15 


184 


12 


fX34 


1X34 


if 


14 


3 


10 X16 


144 


12 


IX 3| 


IX 3f 


2 


1A 


3 


12 X19 


17 


12 


IX .8f 


|X3| 


2 


14 


34 


14 X21 


18f 


12 


IX 44 


1 X44 


2 


U 


34 


15 X22i 


20 


16 


1X44 


1 X4i 


2 


if 


3* 


16 X23£ 


214 


16 


IX 44 


1 X44 


24 


l* 


34 


18 X25 
20 X27i 


22| 
25 


16 


1 X4| 


14X4f 

14X5 




1* 

144 


34 
3| 


20 


1 X5 









PIPING AND PIPE FITTINGS 



615 



TABLE 79. 



SCHEDULE OF STANDARD FLANGES FOR EXTRA HEAVY STEEL PIPE, FITTINGS, 
AND VALVES ADOPTED JUNE 28, 1901, BY VALVE AND FITTING MANUFAC- 
TURERS. SUITABLE FOR PRESSURES FROM 125 TO 250 POUNDS PER SQUARE 
INCH. 





Diameter of 


Thickness of 


Diameter of 




Diameter of 


Size of Pipe, 
Inches. 


Flange, 


Flange, 


Bolt Circle, 


Number of 
Bolts. 


Bolts, 


Inches. 


Inches. 


Inches. 


Inches. 


2 


6i 


* 


5 


4 


1 


H 


7* 


1 


51 


4 


1 


3 


8i 


11 


6f 


8 


f 


3* 


9 


** 


n 


8 


1 


4 


10 


U 


n 


8 


1 


H 


10i 


ift 


8$ 


8 


I 


5 


11 


U 


9i 


8 


• 1 


6 


m 


i* 


lOf 


12 


1 


7 


14 


H 


Hi 


12 


I 


8 


15 


if 


13 


12 


I 


9 


16 


l! 


14 


12 


I 


10 


17J 


if 


m 


16 


I 


12 


20 


2 


171 


16 


I 


14 


22i 


a* 


20 


20 


1 


15 


23i 


2^ 


21 


20 




16 


25 


2* 


22i 


20 




18 


27 


2| 


24i 


24 




20 


29^ 


»* 


26f 


. 24 


1* 


22 


31$ 


2f 


281 


28 


1* 


24 


34 


21 


311 


28 


1* 



TABLE 80. 
LOSS OF HEAT FROM BARE STEAM PIPE.* 

Still Air. 



Authority of 
Test. 



Barrus 

Do 

Do 

Hudson Beare 
"130 lbs.".... 



Jacobus . 
Brill.... 



Descriptive Refer- 
ences. 



i Power, Dec, 1901 
Trans. A.S.M.E.,vol. 
xxin ; Stevens Ind., 
Vol. xix, p. 388. 



Stevens Ind., 

xix, p. 388, 

Stevens Ind., 

XVIII. 

Trans. A.S.M.E., 
Vol. xvi. 



Vol. 
Vol. 



i 

o 


ft, W) 


of 

CO 

I-. 

C be 

.2 O 


s 


1 

55 

p. . 


si 

53 S 


ounds of Steam 
Condensed per 
Sq. Ft. per Hour. 


oq 


OQ 


02 


02 *" 

325 


56 


Q 


iM 


2 


63.57 


82 


268.6 


0.915 


2 


63.92 


149 


365 


63.3 


302.2 


1.150 


10 


98.33 


149 


365 


73.6 


291.7 


1.085 


3.53f 


8.13 


135 


358 


67 


291 


1.050 


2 


50.66 


128 


354 


80.1 


274.6 


0.994 


2 


7.6? 


53 


301 


71.2 


229.6 


0.707 


8 


135.4 


110 


344 


75.5 


269 


0.834 



be 
a> 

P 

02 » g 

t_ u 33 
_• &.S 

£> ft . 



£n Q 



3.01 
3.25 
3.18 
3.10 
3.13 

2.78 

2.71 



* C. P. Paulding, Stevens Indicator, Vol. xix, p. 388. t Outside diameter. 



616 



STEAM POWER PLANT ENGINEERING 



331. Coverings. — Steam pipes, feed-water pipes, boiler steam 
drums, receivers, separators, etc., should be covered with heat-insu- 
lating material to reduce radiation losses to a minimum. For most 
practical purposes the loss of heat from a bare steam pipe or drum 
may be taken as 3 B.T.U. per square foot per hour per degree differ- 
ence in temperature, Table 80. The actual loss depends upon the 
diameter of the pipe, on its position whether vertical or horizontal, the 
nature of the surface, and the velocity of the surrounding air currents. 
For a detailed analysis of these various influences, and interesting 
information on the transmission of heat, the reader is referred to 
Paulding's " Steam in Covered and Bare Pipes." 

By properly applying any good commercial covering, from 75 per 
cent to 90 per cent of the heat loss may be prevented. (See Fig. 
344 and Table 81.) 




Diagram *. Heat 



Fig. 344. Efficiency of Pipe Coverings. 

Example : Required the saving per annum due to covering a pipe 
10 inches in diameter and ltfO feet long; steam pressure 150 pounds; 
average temperature of the air 76 degrees F. ; cost of covering applied 
65 cents per running foot; efficiency of covering 85 per cent; cost of coal 
$2.50 per ton; plant to operate 14 hours per day and 300 days per year. 

The temperature of steam at 150 pounds pressure = 366 degrees F. 



PIPING AND PIPE FITTINGS 



617 



Difference of temperature between the steam and air = 366 — 76 
= 290 degrees F. 

Loss per square foot per hour, bare pipe = 3 X 290 = 870 B.T.U. 

Loss per square foot per day, bare pipe = 870 X 14 = 12,180 B.T.U. 

Loss per square foot per year, bare pipe = 12,180 X 300 = 3,654,000 
B.T.U. 

100 lineal feet of 10-inch pipe has an external surface of 282 square 
feet. Therefore the loss per year from the bare pipe is 

282 X 3,654,000 = 1,030,000,000 B.T.U. (approx.). 



TABLE 81. 

EXPERIMENTS ON STEAM-PIPE COVERINGS. 

("Condensation of Steam in Covered and Bare Pipes " [Paulding].) 



Kind of Covering. 



Hair felt 

Do 

Remanit for interme- 
diate pressure. 
Remanit for high pres- 
sure. 

Mineral wool 

Champion mineral wool 

Rock wool 

Asbestos sponge felted 

Do 

Do 

Magnesia 

Do 

Do 

Do 

Do 

Do 

Asbestos, Navy Brand 

Do 

Do 

Manville sectional. . . 

Do 

Do 

Asbestos air cell 

Do 

Asbestos fire felt 

Do 

Do 

Fossil meal 

Riley cement 



Diam. 


Thick- 


of Test 


ness of 


Pipe, 


Cover- 


Inches. 


ing, 




Inches. 


2 


0.96 


8 


0.82 


2 


0.88 


2 


1.30 


8 


1.30 


8 


1.44 


8 


1.60 


2 


1.125 


10 


1.375 


2 


1.14 


4 


1.12 


2 


1.09 


8 


1.25 


2 


1.08 


2 


1.00 


10 


1.19 


2 


1.20 


2 


1.125 


10 


1.375 


8 


1.70 


2 


1.31 


4 


1.25 


4 


1.12 


2 


0.96 


8 


1.30 


2 


1.00 


2 


0.99 


8 


0.75 


8 


0.75 



Temperatures 
F. 



Steam 



302.8 
348.3 
304.5 

306.6 



344.1 
346.1 
344.1 
364.8 
364.8 
309.2 
388.0 
354.7 
344.1 
310.9 
365.2 
365.2 
309.2 
365.2 
365.2 
345.5 
354.7 
388.0 
388.0 
303.3 
344.7 
354.7 
307.4 
347.1 
347.9 



Air. 



71.4 
69.0 
73.3 

76.1 



58. 
74. 



63.0 
60.7 
62.8 
79.4 
72.0 
80.1 
66.3 
81.6 
64.6 
66.0 
79.4 
64.6 



78 

80 

72 

72 

72 

79 

80 

72.5 

75.3 

74.3 



B.T.U. per 
Hour per 

Square Foot 
of Pipe 
Surface. 



Total. 



89.6 
117.9 
100.3 

83.7 

81.3 

86.1 

72.0 

145.0 

85.0 

59.7 

147.0 

155.8 

106.6 

69.8 

155.0 

103.0 

69.9 

176.0 

112.0 

93.4 

157.0 

143.0 

166.0 

165.5 

133.5 

198.0 

180.0 

238.0 

260.0 



Per 
Degree 
Differ- 
ence. 



0.387 
0.422 
0.434 

0.363 

0.284 

0.317 

0.256 

0.477 

0.248 

0.260 

0.465 

.567 

,384 

304 

,515 

,347 

,304 

.585 

0.375 

0.394 

0.572 

0.453 

0.525 

0.716 

0.502 

0.721 

0.766 

0.876 

0.950 



Date 

of 
Test 



1901 
1894 
1901 

1901 

1894 
1894 
1894 
1901 
1901 
1901 
1896 
1896 
1895 
1901 
1901 
1901 
1901 
1901 
1901 
1894 
1896 
1896 
1896 
1901 
1894 
1896 
1901 
1894 
1894 



Testing Ex- 
pert. 



Jacobus 

Brill 

Jacobus 

Jacobus 

Brill 

Brill 

Brill 

Barrus 

Barrus 

Jacobus 

Norton 

Paulding 

Brill 

Jacobus 

Barrus 

Barrus 

Jacobus 

Barrus 

Barrus 

Brill 

Paulding 

Norton 

Norton 

Jacobus 

Brill 

Paulding 

Jacobus 

Brill 

Brill 



618 STEAM POWER PLANT ENGINEERING 

Assuming a net available heat value of 10,000 B.T.U. per pound for 
the coal, the equivalent coal consumption is 51.5 tons, valued at 
51.5 X $2.50 = $128.75. 

The covering will save 85 per cent of this, or $109.50 per annum. 

The pipe covering applied will cost 100 X 0.65 = $65.00. 

In this case the covering will pay for itself in considerably less than 
a year. 

Pipe covering is applied in sections molded to the required form 
and held to the pipe by bands, or may be applied in a plastic form. 
The former is more readily applied and removed, and is usually 
adopted for pipes, while the valves and fittings are sometimes covered 
with plastic material. Piping should be tested under pressure before 
being covered, since leaks destroy the efficiency and life of the cover- 
ing. If the surrounding atmosphere is moist the covering should be 
given two or three coats of good paint. Coverings are sometimes 
applied to cold-water pipe to prevent sweating in a humid atmosphere. 

Pipe Coverings: Power, July, 1904, p. 407, Aug., 1904, p. 482, May, 1903, p. 239, 
Dec, 1901, p. 32; St. Ry. Jour., Nov. 29, 1902, p. 875; Engr., Lond., May 27, 1904, 
p. 547; Eng. Review, Nov., 1898, p. 15; Am. Elecn., May, 1903; Engng., Aug. 7, 1903; 
Mech. Engr., Nov. 25, 1905; Elec. World and Engr., April 6, 1901; Stevens Ind., 
Oct., 1902; Trans. A.S.M.E., 16-827, 23-791. 

Identification of Power House Piping by Colors: Power & Engr., Apr. 26, 1910, p. 752. 

332. Expansion. — One of the most difficult problems in the design 
of a piping system is the proper provision for expansion and contrac- 
tion due to change in temperature. If a pipe is immovably fixed at 
both ends and under no strain when cold, and the temperature is 
increased, as by the admission of steam, it is subjected to a com- 
pression proportional to the rise in temperature (within the elastic 
limit). For example, a 6-inch standard extra heavy wrought-iron 
pipe 200 feet long at 66 degrees F., if heated to 366 degrees F. (the 
temperature corresponding to steam at 165 pounds per square inch 
absolute pressure), will exert an axial force of 

P = EA (t t — t)ft. (Mechanics of Engng., Church, p. 218.) (147) 

P = force in pounds. 

E = modulus of elasticity, 30,000,000. 

t t = final temperature, degrees F. 
t = initial temperature. 

jt* = coefficient of expansion, 0.0000075. 

A = sectional area of the pipe material, 8.5 square inches. 

Hence 

P = 30,000,000 X 8.5 (366 - 66) 0.0000075. 
= 573,750 pounds. 



PIPING AND PIPE FITTINGS 619 

Unless well braced throughout its entire length the pipe will buckle 
and become distorted. If free to expand its length would increase. 
The temperature of the pipe is always less than that of the steam on 
account of radiation from the outer surface and varies with the effi- 
ciency of the covering. But ignoring radiation the increase in 
length is 

1 = n(t x -t) L, (148) 

in which 

I = increase in length, inches. 
L = length of pipe, inches. 

Other notations as in (147). 

Substituting in (148), t t = 366. 
t = 66. 
Pl = 0.0000075. 
L = 2400. 

I = 0.0000075 (366 - 66) 2400 
= 5.4 inches. 

Since the forces produced by expansion are practically irresistible, 
the pipe is invariably allowed to expand freely by suitable means so 
as not to strain the connections. The coefficients of expansion per 
degree difference in temperature for various pipe materials are given 
in Table 82. 

Headers less than 50 feet in length usually require no special pro- 
visions for expansion provided the ends are free and the leads to and 




a b «^^ c o 

Fig. 345. Types of Expansion Pipe Bends. 

from the header are not too short, the pipe usually being anchored at 
the middle and permitted to expand in either direction. Free expan- 
sion of the feeders may be provided for 

(1) By long radius bends, as in Fig. 345. 

(2) By double swing screwed fittings, as in Fig. 346, or 

(3) By packed expansion joints, Fig. 347. 

Where practicable the long radius bends will prove most satis- 
factory. The radius of the bend should not be less than 5 diameters 
of the pipe, and larger if possible. The length of straight pipe at the 



620 



STEAM POWER PLANT ENGINEERING 



end of each bend should not be less than twice the diameter of the 
pipe measured from the face of the flange. 

On account of the great strains to which the joints of pipe bends are 
subjected, the welded joint, G, Fig. 343, is recommended as giving the 
best results. The next best is the lap joint, F, Fig. 343. 

TABLE 82. 

COEFFICIENTS OF LINEAR EXPANSION PIPING MATERIALS. 



Material. 



Wrought iron and mild steel. . . 

Wrought iron 

Cast iron 

Cast steel 

Hardened steel . . . 

Nickel-steel, 36 per cent Nickel 

Copper, cast 

Copper, wrought 

Lead 

Cast brass 

Brass wire and sheets 

Tin cast 

Tin hammered 

Zinc cast 

Zinc hammered 



Temperature 
Range. 


Mean Coeffi- 
cient per De- 
gree F. 


32-212 


0.00000656 


32-572 


0.00000895 


32-212 


0.00000618 


32-212 


0.00000600 


32-212 


0.00000689 


32-572 


0.00000030 


32-212 


0.00000955 


32-572 


0.00001092 


32-212 


0.00001580 


32-212 


0.00001043 


32-212 


0. 00001075 


32-212 


0.00001207 


32-212 


0.00001500 


32-212 


0.00001633 


32-212 


0.00001722 



LINEAR EXPANSION OR CONTRACTION OF CAST IRON IN INCHES PER 
100 FEET, — DEGREES F. 



Temperature Difference. 


Expansion. 


Temperature Difference. 


Expansion. 


100 
150 
200 
250 


0.72 
1.1016 
1.5024 
1.9260 


300 
400 
500 
600 
800 


2.376 
3.360 
4.440 
5.616 

7.872 









Multiply by 1 .1 for wrought mild steel. 
Multiply by 1 .5 for wrought copper. 
Multiply by 1.6 for wrought brass. 

Fig. 345, A, B, C, D shows applications of pipe bends to straight pipe 
runs. A is the cheapest and most common arrangement for all sizes of 
pipe. B is a modification for limited center to center spaces. C shows 
a common method of taking up expansion in straight runs of pipe of 
very large diameters where the space requirements prohibit the use of 






PIPING AND PIPE FITTINGS 



621 



a single U bend. Here the main runs are connected to manifolds which 
in turn are connected by a number of small U bends, the equivalent 
areas of which correspond to that of the large pipes. This makes a more 
flexible connection than if a single U bend were used. The arrange- 
ment D does away with the elbows required in A, 
but is not applicable to pipes over 8 inches in 
diameter. 

Figs. 357 and 358 show applications of pipe 
bends to boiler and header connections. 

Fig. 346 shows a double swing screwed joint 
in which expansion causes the fittings to turn 
slightly and 
thus relieve the 
strain. This 
method is usu- 
ally adopted 
where long ra- 
dius bends are 
not practicable 
on account of 
lack of space and where screwed fittings are used. 

Slip joints, Fig. 347, are now little used except with very large pipes 
and where space prohibits long radius bends. When slip joints are 
employed the pipe must be securely anchored to prevent the steam 
pressure from forcing the joint apart and at the same time permit the 
pipe in expanding to work freely in the stuffing box. Sagging of the 
pipe on either side, which might cause binding in the joint, is prevented 
by suitable supports. 

Expansion in Steam Pipes: Power, July, 1906, p. 426, Jan., 1904, p. 30, March, 
1904, p. 160, Oct., 1904, p. 609, Dec, 1900; Am. Elecn., 10-432; Engr., U.S., 
Feb. 1, 1904, p. 125; Eng. News, 44-194, 47-468, 50-487; Power, June 2, 1908. 




FRONT ELEVATION 

Fig. 346. "Double-Swing 
Expansion Joint. 




Fig. 347. Slip Expansion Joint. 



333. Pipe Supports and Anchors. — Pipe lines must be supported to 
guard against excessive deflection and vibration. Supports are con- 
veniently classified as (1) hangers, (2) wall brackets, and (3) floor stands. 

Fig. 348 illustrates a type of hanger for suspending pipes from I 
beams. The supports being free to swing, no provision for expansion 
is necessary. A properly designed hanger may be readily removed 
without disturbing the pipe line, and should be adjustable to facilitate 
" lining up." If of rigid construction the lower end should be provided 
with a roller. 

Fig. 349 gives the details of a wall bracket with rolls and roll binder. 



622 



STEAM POWER PLANT ENGINEERING 



Supports adjacent to long radius bends should be provided with roll 
binders as illustrated to prevent the pipe from springing laterally, but 






Fig. 348. A Typ- 
ical Pipe Hanger. 



Fig. 349. A Typical Wall 
Bracket with Binding Roll. 



Fig. 349a. A Typ- 
ical Floor Stand. 



they may otherwise be omitted. The rollers are often made adjustable 

to facilitate lining up. 

Fig. 349a illustrates a typical floor stand. 
Pipe lines are usually securely anchored 
at suitable points in a manner similar to 
that illustrated in Fig. 350, the pipe rest- 
ing on a saddle and being rigidly clamped 
to the bracket by a flat iron band with 
ends threaded and bolted. This limits 
expansion to one direction and prevents 
excessive strain on the fittings. 

334. General Arrangement of High- 
Pressure Steam Piping. — The general ar- 
rangement of piping depends in a great 
measure upon the space available for en- 
gines and boilers. 

The engine and boiler room may be 
placed 
(1) Back to back, Fig. 361, (2) End to end, Fig. 351, 

(3) Double decked, Fig. 356. 




Fig 350. A Typical Pipe Anchor. 






PIPING AND PIPE FITTINGS 



623 



The back to back arrangement is the most common and, other things 
permitting, is to be preferred on 
account of the short and direct 
connection between engines and 
boilers and the ease of enlarge- 
ment. The engine and boiler 
rooms are separated by a wall, 
and as much of the piping as 
possible is located in the boiler 
room. 

The end to end arrangement 
is ordinarily limited to situa- 
tions where the distribution of 
space precludes the back to back 
system. 

The double decked arrangement 
is frequently used where ground 
space is limited or expensive. 

Engines and boilers are con- 
nected in a variety of ways 
through steam headers as shown 
in the following examples : 

(1) Single header, Fig. 361, 

(2) Duplicate headers, Fig. 
352, 

(3 ) Loop or ring header, Fig. 
353, 

(4) The "unit" system, Fig. 
357. 

The single header system is per- 
haps the most common, since it 
embodies simplicity, low first 
cost, and provision for extension. 

The duplicate system is losing 
favor, since experience shows 
that the extra cost of the du- 
plicate mains will usually give 
better returns in continuity of 
operation and maintenance if in- 
vested in high-grade fittings on 
a single pipe system. 

The loop header is well adapted for the power plants of tall office build- 




DZ4 



STEAM POWER PLANT ENGINEERING 




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1 








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Fig. 352. Typical Duplicate Header System. 




VALVE ©: ' 



rt 



--) 



Fig. 353. Typical Loop Header. 



PIPING AND PIPE FITTINGS 



625 



ings, Fig. 355, in which a large number of steam engines, elevator pumps, 
air compressors, and miscellaneous steam-consuming appliances are 
crowded together in a comparatively small space. 

Large modern power plants are, by the latest practice, divided into 
complete and independent units, as in Fig. 440, each prime mover 
having its own boiler equipment, coal and ash-handling machinery, 
feed pumps, and piping, operated independently of the rest of the plant, 
though provision is made whereby any boiler equipment may provide 
steam for any prime mover. 




D 



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li 



rm=° 



; wgr 



Fig. 354. Typical By-Pass System. 



The power plant of the Manhattan Elevated Railway Company, 
New York, is practically divided into eight sections each consisting 
of an engine and eight boilers, the boilers being " double decked " 
(Fig. 356). 

The branch pipes from the upper and lower batteries lead into 18- 
inch headers, the steam from each being conducted to a receiver reservoir 
36 inches in diameter and 20 feet long in the engine room basement 
directly behind each engine, from which the two high-pressure cylinders 
are supplied. Gate valves are used in each boiler branch, one close to 
the boiler and another near the header, and also in the steam pipes 
near the reservoir. The steam headers for each of the eight units are 
connected by a main which equalizes the pressure and allows a deficiency 
in one unit to be made up from the others. 

Figs. 357 and 358 show the general arrangement of the steam piping 
at the Yonkers power house of the New York Central. The turbines 



626 



STEAM POWER PLANT ENGINEERING 




PIPING AND PIPE FITTINGS 



627 




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628 



STEAM POWER PLANT ENGINEERING 



a qnp a .. ( 




Detail Plan 



(Power) 



Fig. 358. Details of Boiler Steam Piping, Yonkers Power House of the New York 

Central R.R. 



PIPING AND PIPE FITTINGS 



629 



are connected in pairs by 14-inch loops, each turbine taking steam 
from either of two banks of four boilers. The high-pressure steam 
piping is of mild steel with modified reenforced "Van Stone" joints. 
The high-pressure valves are of the split disk pattern with semi-steel 
bodies. Expansion is taken up by the long sweep bends. 




Fig. 359. Overhead Boiler Piping, Quincy Point Power Plant of the Old Colony St. 
Ry. Co., Quincy Point, Mass. 



Plants using superheated steam are ordinarily piped to supply 
saturated steam to the auxiliaries as illustrated in Fig. 359. The 
boiler branch E, leading to the main header, normally supplies super- 
heated steam to the engines. C is an auxiliary main supplying the 
air pumps, stoker engines, and other auxiliaries with saturated steam 
from branch pipe D. 

Steam Piping: Power, Feb. 23, 1909, Nov., 1905, p. 683, April, 1904, p. 213, Feb., 
1904, p. 90, Sept., 1904, p. 540, Nov., 1904, p. 677, July, 1903, p. 356; Ir. Tr. Rev., 
May 29, 1902; Ir. Age, Dec. 11, 1902; Met. Work, Feb. 14, 1903; St. Ry. Rev., Nov. 
20, 1904, p. 869; St. Ry. Jour., Oct. 15, 1904, p. 441 ; Am. Elecn., March, 1905, p. 127; 
Elec. Engr., Lond., Dec. 23, 1904; Engr. U.S., Feb. 15, 1905, Dec. 1, 1904; Elecn., 
Lond., July 21, 1899; Eng. Rec, 48-90; Trans. A.S.M.E., 15-536; Cass. Mag., June, 
1906. 

335. Main Steam Headers. — Until quite recently it was the usual 
practice to make the area of the steam header equivalent to the com- 
bined areas of the feeders, but the function of the header is now 
regarded as that of an equalizer rather than a storage reservoir. In 
the various large power houses recently built in New York City, with 
ultimate capacities of from 60,000 to 150,000 kilowatts, the largest 
steam headers are not over 16 inches in diameter. In some recent 



630 



STEAM POWER PLANT ENGINEERING 



designs the pipes leading from the header to the engines are two sizes 
smaller than called for by the engine builders. In this case large 
receiver separators two to four times the volume of the high-pressure 
cylinder are provided near the throttle as in Fig. 356. The pipes 
between receiver and engine are full size. The object of the arrange- 
ment is to give (1) a constant flow of steam, (2) a full supply of steam 
close to the throttle, and (3) a cushion near the engine for absorbing 
the shock caused by cut-off. With moderately superheated steam and 
boiler pressures from 125 to 150 pounds a velocity of 8000 feet per 
minute is allowed in the header and as high as 9000 feet per minute 
between header and receiver. With steam turbines velocities as high 
as 12,000 feet per minute are permissible, provided the pipe is less than 
50 feet in length and practically free from sharp bends. Main headers 
are ordinarily constructed of mild steel, though cast-iron and cast- 
steel headers are not uncommon. Cast headers permit of fewer joints 
and are well adapted to situations where a number of branches are 
closely grouped as in Fig. 361. Cast-iron headers are employed in the 
Manhattan Elevated Railway power station, New York. 



<fi> 



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»oi <r- — if 




z ^^^ ^ ^^^^^^^^^^^^ ^ ^^ ^ 



Nffp@a@ofp -y- mm~m 




Fig. 360. Steam Header and Branches, Grand Rapids, Grand Haven and Muskegon 

Ry. Co. Power House. 



The proper arrangement and number of valves in the main header 
and feeders has been a subject of much consideration. Figs. 356 to 
361 show some of the different successful arrangements in recent 
installations. In Fig. 360 there is but one valve between the boiler 
nozzle and the main header, while in Fig. 361 there are two. The 
latter is the more common arrangement. Where two valves are 



PIPING AND PIPE FITTINGS 



631 




o 

'£ 
a 
o 



632 STEAM POWER PLANT ENGINEERING 

placed in a feeder they should be arranged so as not to form a pocket 
for the accumulation of leakage. In a number of recent installa- 
tions, Fig. 358, the valve nearest the boiler is* of the " automatic stop 
and check " type, its function being the automatic cutting off of the 
steam from the header should the pressure in the boiler suddenly drop 
as in case of blowing out a tube. 

Arrangement of Steam Piping: Am. Elecn., April, 1905, June, 1902, p. 257, June, 
1900; Engr. U.S., Dec. 1, 1904; Mech. Engr., Nov. 4, 1905; Power, Sept., 1904, p. 511, 
July, 1902; Eng. News, Nov. 26, 1903, p. 487; Elec. Rev., Lond., Aug. 11, 1899, 
p. 251; St. Ry. Rev., Jan., 1900, p. 12; Nov., 1904, p. 869. 

336. Flow of Steam in Pipes.*: — The several accepted formulas 
relating to the flow of steam in pipes have been based upon a few 
experiments limited to pipes of small diameter, hence the application 
of these formulas to larger pipes or to conditions other than those 
under which they were deduced is apt to lead to considerable error. 
In small plants extreme accuracy in determining the proper sizes is 
not necessary; it is better to err in the installation of too large a pipe 
than one too small. In larger stations, however, where the pipes are 
large and the pressure high, the cost of the piping increases very 
rapidly with the size. For example, the cost of 10-inch high-pressure 
fittings is from 15 to 20 per cent greater than 9-inch fittings, and in 
large installations this first cost item may be of considerable impor- 
tance. 

The simplest and most commonly used formula is based upon an 
allowable steam velocity of 6000 feet per minute, friction and other 
causes of drop in pressure being disregarded; thus, for a velocity of 
6000 feet per minute, 

d = 0.175 



(*)*. a«> 



in which 

d 4= diameter of the pipe in inches, 

y = density of the steam in pounds per cubic feet, and 

W == weight of steam flowing in pounds per minute. 

In determining the diameter of the steam pipe opening for recipro- 
cating engines a much lower velocity than 6000 feet per minute is 
assumed, to allow for the various conditions of operation. Average 
practice gives the constant in equation (149) a value of 0.3 instead of 
0.175 when used in this connection. 

Equation (149) gives satisfactory results for pipes under 100 feet in 
length and between 4 and 8 inches in diameter; for larger diameters 
the velocity could be increased with advantage; for smaller diameters 

* See author's original paper, Power, June, 1907, .p. 377. 



PIPING AND PIPE FITTINGS 



633 



or greater lengths friction and condensation would cause considerable 
drop in pressure and some one of the approved formulas in Table 83 
should be used instead. 




1000 2000 3000 4000 5000 6000 SCC0 10000 12000 14000 

Mean Velocity, Feet per Minute 



16000 13000 



Fig. 362. Drop in Pressure for Various Velocities and Pipe Sizes. Initial Pressure 100 
Pounds Gauge, Length of Pipe 100 feet. 

A large drop in pressure means a small pipe and high velocity with 
consequent decrease in condensation, but a point is soon reached 
where the economy in the size of pipe is more than offset by the loss 
in friction. There seems to be no fixed rule for determining the drop 



634 



STEAM POWER PLANT ENGINEERING 



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636 STEAM POWER PLANT ENGINEERING 

most suitable for any given set of conditions. In current practice the 
drop in pressure between boiler and engine ranges from a fraction of 
one pound to four pounds per square inch per 100 feet of pipe, with an 
average between one and two pounds. 

Table 83 gives a few of the best known formulas for the flow of 
steam, and Table 84 a comparison between them with respect to 
velocity, weight discharged, diameter, and the drop in pressure. 

Formula 11, Table 83, is the most commonly accepted, and the 
curves in Fig. 362 are based upon it, assuming a steam pressure of 100 
pounds absolute and pipe lengths of 100 feet. Within the limit of 
12,000 feet per minute velocity and 10 pounds per square inch drop in 
pressure the curves are sufficiently accurate for all practical purposes, 
but beyond this range the results are purely conjectural and may not 
be accurate, as no recorded experiments have been conducted at these 
high velocities or with pipes of large diameters. 

Though applicable directly to pipes 100 feet long with mean pres- 
sure of 100 pounds per square inch absolute, they may be used for any 
length or pressure. For example, for any length other than 100 feet, 
multiply the drop given in the curves by the required length in feet 
and divide by 100. For any pressure other than 100 pounds abso- 
lute, multiply the drop given in the curves by 0.2271 (density of steam 
in pounds) and divide by the density of steam at the required 
pressure. 

Table 85 is the table ordinarily used in connection with the flow of 
steam and is calculated from equation 11. Table 86 is based upon 
equations 4 to 12. The results differ slightly from those in Table 85, 
though the latter is more comprehensive. The left-hand half of Table 
86 gives the discharge in pounds per minute for pipes of various 
diameters corresponding to drop of pressure as given on the right-hand 
side in the same horizontal line; e.g., a 6-inch pipe 100 feet long dis- 
charges 371 pounds of steam per minute for a drop of 16.4 pounds at 
100 pounds pressure. 

337. Equation of Pipes. — It is frequently desirable to know what 
number of one sized pipes will be equal in capacity to another pipe. 

According to the formulas in Group II, Table 84, the weights dis- 
charged vary with the square root of the fifth power of the diameter; 
that is, the number of pipes equal in capacity to any given pipe may 
be determined from the equation 

N 1 = d^-^dX (150) 

in which N x = number of pipes of diameter d l equal in capacity to a 
pipe of diameter d; d, and d in inches. 



PIPING AND PIPE FITTINGS 



637 



TABLE 85. 
FLOW OF STEAM THROUGH PIPES (BABCOCK). 





Diameter of Pipe, in Inches. Length of each = 240 diameters. 


















sure by 
















Gauge. 


i 


1 


l* 


2 


2* 


3 


4 


Pounds per 
















Square Inch. 


















Weight of Steam per Minute, in pounds, with One Pound Loss of Pressure. 


1 


1.16 


2.07 


5.7 


10.27 


15.45 


25.38 


46.85 


10 


1.44 


2.57 


7.1 


12.72 


19.15 


31.45 


58.05 


20 


1.70 


3.02 


8.3 


14.94 


22.49 


36.94 


68.20 


30 


1.91 


3.40 


9.4 


16.84 


25.35 


41.63 


76.84 


40 


2.10 


3.74 


10.3 


18.51 


27.87 


45.77 


84.49 


50 


2.27 


4.04 


11.2 


20.01 


30.13 


49.48 


91.34 


60 


2.43 


4.32 


11.9 


21.38 


32.19 


52.87 


97.60 


70 


2.57 


4.58 


12.6 


. 22.65 


34.10 


56.00 


103.37 


80 


2.71 


4.82 


13.3 


23.82 


35.87 


58.91 


108.74 


90 


2.83 


5.04 


13.9 


24.92 


37.52 


61.62 


113.74 


100 


2.95 


5.25 


14.5 


25.96 


39.07 


64.18 


118.47 


120 


3.16 


5.63 


15.5 


27.85 


41.93 


68.87 


127.12 


150 


3.45 


6.14 


17.0 


30.37 


45.72 


75.09 


138.61 



Initial 

sure by Gauge 

Pounds per 

Square Inch. 



1 

10 

20 

30 

40 

50 

60 

70 

80 

90 

100 

120 

150 



Diameter of Pipe, in Inches. Length of each =240 diameters. 



10 



12 



15 



18 



Weight of Steam per Minute, in Pounds, with One Pound Loss of Pressure. 



77.3 
95.8 
112.6 
126.9 
139.5 
150.8 
161.1 
170.7 
179.5 
187.8 
195.6 
209.9 
228.8 



115.9 
143.6 
168.7 
190.1 
209.0 
226.0 
241.5 
255.8 
269.0 
281.4 
293.1 
314.5 
343.0 



211.4 
262.0 
307.8 
346.8 
381.3 
412.2 
440.5 
466.5 
490.7 
513.3 
534.6 
573.7 
625.5 



341.1 
422.7 
496.5 
559.5 
615.3 
665.0 
710.6 
752.7 
791.7 
828.1 
862.6 
925.6 
1009.2 



502.4 

622.5 

731.3 

824.1 

906.0 

979.5 

1046.7 

1108.5 

1166.1 

1219.8 

1270.1 

1363.3 

1486.5 



804 
996 
1170 
1318 
1450 
1567 
1675 
1774 
1866 
1951 
2032 
2181 
2378 



1177 
1458 
1713 
1930 
2122 
2294 
2451 
2596 
2731 
2856 
2975 
3193 
3481 



For any other length divide 240 by the given length expressed in diameters and multiply 
the tabular quantity by the square root of this quotient, which will give the flow for one pound 
loss of pressure. Conversely, dividing the given length by 240 will give the loss of pressure 
for the flow given in the table. 



638 



STEAM POWER PLANT ENGINEERING 



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PIPING AND PIPE FITTINGS 639 

According to the formulas in Group I, Table 84, the weights dis- 
charged vary as \ d 5 -s- 1 1 + —j- ) \ and the equation becomes 

*• " TTIe * "Tie (151 > 

I d + dj 

= /J(d,+3.6)\* 

U' W + 3.6)/ (15i) 

- J^*™- d53) 

From (150) and (153) we see that the values of N t are practically 
the same for either equation when the ratio of d to d ± is small and that 
they differ widely for large ratios. For example, according to (150), 
5.7 eight-inch pipes are equivalent in capacity to one sixteen-inch pipe, 
whereas (152) gives 6.15. The difference is negligible. Again, 
according to (150), 180 two-inch pipes are equivalent in capacity to 
one sixteen-inch pipe, whereas (153) gives 274. The difference is con- 
siderable. Equation (153) is most commonly accepted and is the basis 
of Table 87. 

338. Friction through Valves and Fittings. — The formulas out- 
lined in Table 83 are strictly applicable only to well-lagged pipes, free 
from bends or obstructions of any kind such as valves or fittings, 
which greatly increase the resistance of the flow of steam. If these 
obstructions must be considered, it is customary to allow for them by 
assuming an added length of straight pipe equivalent in resistance 
to the various fittings and bends. Unfortunately, the few tests which 
have been made for the purpose of determining the resistance of vari- 
ous pipe fittings give discordant results, and in the absence of more 
recent data the rules given by Briggs (" Warming Buildings by Steam") 
are probably as accurate as any. 

According to Briggs, the length of pipe in inches equivalent to the 
resistance of one standard 90-degree elbow is 

L = 76 d -*- (l + y) < 154 > 

and to that of one globe valve 

L = 114 d + (l + y)' < 155 > 

The resistance of gate valves is not considered. 



640 



STEAM POWER PLANT ENGINEERING 



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I 2 






PIPING AND PIPE FITTINGS 



641 




642 STEAM POWER PLANT ENGINEERING 

339. Exhaust Piping, Condensing Plants. — The exhaust piping in 
condensing plants is arranged either according to (1) the independent 
or (2) the central condensing system. In the former each engine is 
provided with an independent condenser and air pump. In case the 
vacuum " drops " or it is desired to operate non-condensing, the steam 
is discharged through a branch pipe with relief valve to the atmos- 
phere, Figs. 3 and 219. When there are a number of engines in one 
installation the atmospheric pipes lead to a common free exhaust main, 
which, on account of its great size, is ordinarily constructed of light- 
weight riveted steel pipe. The short connection between engine and 
condenser is usually made with lap-welded steel pipe, since riveted 
joints are apt to leak, due to the engine vibrations. In a central con- 
densing plant, Fig. 226, the several engines exhaust through a com- 
mon main into a single large condenser. An atmospheric relief valve 
is usually provided in connection with the condenser, and no free 
exhaust main is necessary. Several arrangements of condenser piping 
are illustrated in Figs. 219 to 228. 

340. Exhaust Piping, Non-Condensing Plant. — Webster Vacuum 
System. In the majority of non-condensing plants the exhaust steam 
is used for heating purposes. One of the best-known systems of 
exhaust steam heating, in which the back pressure on the engine is 
reduced by circulating below atmospheric pressure, is that known as 
the Webster combination system. The general arrangement is illus- 
trated in Fig. 2 and the principles of operation are described in para- 
graph 3. It has the advantage of affording (1) minimum back 
pressure on the engine; (2) effective and continuous drainage of 
condensation from supply pipes and radiators; (3) continuous 
removal of air and entrained moisture from confined spaces; (4) inde- 
pendent regulation of temperature in each radiator; (5) continu- 
ous return of condensation to the boiler; (6) utilization of part of 
the exhaust steam for preheating the feed water; and (7) automatic 
regulation. Fig. 363 gives a diagrammatic arrangement of the piping 
and appurtenances in a typical installation. The characteristic feature 
of this system is the automatic outlet valve attached to each part 
requiring drainage, which permits both the water of condensation and 
the non-condensable gases to be removed continuously. The radiator 
temperature may be regulated by Varying the quantity of steam sup- 
plied either by hand, or automatically by thermostatic control. The 
Webster valve, Fig. 364, enables the vacuum to withdraw the water of 
condensation as fast as it is formed irrespective of the pressure in the 
radiator, hence the supply may be throttled to such an extent that the 
temperature in the radiator is practically as low as that of steam cor- 



PIPING AND PIPE FITTINGS 



643 



responding to the pressure in the vacuum line. The small annular 
space between the inner tube of the float F and the guide H permits 
of a vacuum in the body of the valve. When the water from the 
radiator lifts the float the water is drawn into the returns pipe. The 




S^Wi^!^ 




OUTLET 

Fig. 364. Webster Air Valve. 



Fig. 365. Automatic Vacuum 
Valve, Illinois Engineering Co; 



valve then returns to its seat and the escape of steam is prevented 
except such as finds its way through the annular space around the 
guide stem H. An improvement on this valve which prevents the 
escape of steam is illustrated in Fig. 365. When steam is admitted to 
the radiator the condensation flows into the valve, righting the float A 
and sealing the outlet B against the passage of steam; as the valve 
fills with water the buoyancy of the float raises it from its seat and 
permits the water to be drawn out; the float falls and reseats on the 
nipple when about a half-inch of water remains in the valve, thus 
maintaining a water seal. 

Screen D prevents scale and dirt from entering the valve proper. 
By-pass H is for emergency use in draining off accumulated water in 
the radiator in case the valve becomes stopped up ; and permits the 
bonnet to be removed without trouble from the accumulated water. 

341. Exhaust Piping, Non-Condensing Plants. — Paul Heating System. 
The Paul vacuum system differs from the Webster in that the con- 
densation, and the air and non-condensable gases are separately 
handled. Referring to Fig. 366, which gives a diagrammatic arrange- 
ment of the piping, the condensed steam gravitates to the automatic 
return tank and pump and is pumped either directly to the boiler or 
through the heater to the boiler. Air and vapor are withdrawn from 
the upper part of the radiator by the Paul exhauster or ejector, E, 
and discharged into the returns tank, which is vented to the atmos- 
phere for the escape of the non-condensable gases. The exhauster 



644 



STEAM POWER PLANT ENGINEERING 




PIPING AND PIPE FITTINGS 



645 



receives its supply of steam through pipe 0, Fig. 367, which shows the 
general arrangement of this apparatus. The piping is in duplicate to 
guard against failure to operate. The suction side of the exhauster is 
connected with the air pipes 

sur.TinN SUCTION 

A, A, Fig. 366. Fig. 368 
gives a section through the 
Paul air or vacuum valve 
which prevents steam from 
blowing into the air pipes 
and permits only air to 
pass. In Fig. 366 the heat- 
ing system is piped on what 
is known as the " one-pipe 
down-feed system;" i.e., 
the exhaust steam is first 
conducted to a distributing 
header in the attic, from 
which the various supply 
pipes are led to the radia- 
tors. The water of conden- 
sation returns through these 
same pipes and gravitates 
to the returns pump. Both 
the supply steam and the 

condensation flow in the same direction. This system is also piped on 
the " one-pipe up-feed," the " two-pipe up-feed," and the " two-pipe 
down-feed " principle. The " one-pipe up-feed " differs from the sys- 
tem just described in that the 
steam flows upward through 
the risers and does away with 
the attic piping. The returns, 
however, flow against the current 
of steam and water hammer 
is more likely to occur than 
with the down-feed system. In 
the two-pipe systems the steam 
supply pipes or risers conduct 
steam only, and the returns 
carry the condensation. The 
one-pipe down-feed is cheaper and simpler and practically as efficient 
as the two-pipe system under normal conditions. It is objection- 
able, however, due to the difficulty of draining the radiator with 




STOP VALVE 

o 



STEAM SUPPLY 



Fig. 367. Paul Exhauster. 



COMPOSITION 




FROM RADIATOR 



TO EXHAUSTER 



Fig. 368. Paul Vacuum Valve. 



646 



STEAM POWER PLANT ENGINEERING 



SSWSSSWASS 



closely throttled supply valve, since the velocity of the entering 
steam prevents the water from returning through the same orifice. 
342. Automatic Temperature Control. — Experience shows that a 
considerable saving in fuel may be effected in the heating plants of tall 
office buildings and similar plants by automatically controlling the 
temperature. Hano)-controlled valves are usually left wide open, and 
when the room becomes too hot the temperature is frequently lowered 
by opening the window, resulting in a waste of heat which may be 
considerable in modern buildings with hundreds of offices. Many 
successful methods of automatic temperature control are available, 
the usual system consisting of thermostats which control the supply of 
heat by means of diaphragm valves, the latter taking the place of the 
usual radiator supply valve. 

Fig. 369 shows a Powers thermostat. The expansible disk U con- 
tains a volatile liquid having a boiling point of about 50 degrees F. 

The pressure of the vapor within the disk 
at a temperature of 70 degrees amounts 
to six pounds to the square inch, and 
varies with every change of temperature, 
causing a variation in the thickness of the 
disk. The disk is attached by a single 
screw to the lever Q, which rests upon 
the screw F as a fulcrum. The flat spring R 
holds the lever and disk against the mova- 
ble flange M . Connecting with the cham- 
ber N are two air passages H and /. The 
thermostat is attached by means of two 
screws at the upper end to a wall plate per- 
manently secured to the wall. This wall 
plate has ports registering with H and I, 
one for supplying air under pressure and the 
other for conducting it to the diaphragm 
motor which operates the valve or damper. 
Air is admitted through H under a pressure 
of about fifteen pounds per square inch, and 
its passage into chamber N is regulated by 
the valve J, which is normally held to its 
seat by a coil spring under cap P. K is an 
elastic diaphragm carrying the flange M, with 
escape valve passage covered by the point of valve L. Valve L tends 
to remain open by reason of the spring. When the temperature rises 
sufficiently expansion of the disk U first causes the valve to seat, its 




Fig. 369 Section Through 
Powers Thermostat. 



PIPING AND PIPE FITTINGS 



647 



spring being weaker than that above valve J. If the expansive 
motion is continued, valve J is lifted from its seat and compressed air 
flows into chamber N, exerting a pressure upon the elastic diaphragm K 
in opposition to the expansive force of the disk. If the tempera- 
ture falls, the disk contracts and the overbalancing air pressure in 
N results in a reverse movement of the flange M, permitting the escape 
valve to open and discharge a portion of the air; thus the air pressure 
is maintained always in direct propor- 
tion to the expansive power (and tem- 
perature) of the disk U. The passage / 
communicates with a diaphragm valve, 
Fig. 370. The compressed air operates 
the diaphragm against a coiled spring 
resistance, so that the movement is 
proportional to the air pressure and 
the supply of steam controlled accord- 
ingly. The adjusting screw G, squared 
to receive a key, carries an indicator by 
means of which the thermostat can be 
set to carry any desired temperature 
within its range, usually from 60 to 80 
degrees. In changing the temperature 
adjustment lever Q forces the disk U 
closer to or farther from the flange M . 

In connecting up the system com- 
pressed air is carried to the thermostat and diaphragm valves, from 
a reservoir through small concealed pipes. 

In the indirect system of heating the dampers are of the diaphragm 
type and the method of regulation is the same as with the direct system. 

343. Feed- Water Piping. — The simplest arrangement of feed- 
water piping may be found in non-condensing plants, in which the 
feed water is obtained under a slight head, such as is afforded by the 
average city supply, and is heated in an open heater by the exhaust 
steam from the engine to a temperature varying from 180 to 210 
degrees F. The hot feed water gravitates from the heater to the 
pump and then is forced to the boiler, or to the economizer if one is 
used. If a meter is used it is generally placed on the discharge side 
of the pump, and should be by-passed to permit it to be cut out for 
repairs. (Fig. 371.) Plants operating continuously should have feed 
pumps in duplicate. In some cases the returns from the heating 
system gravitate to the heater and only enough cold water is added to 
make up the loss from leakage, etc. In other cases the returns gravi- 




Fig. 370 



Diaphragm 



648 



STEAM POWER PLANT ENGINEERING 



tate to a special " returns tank/' from which they are pumped directly 
to the boiler without further heating. Occasionally a live-steam 
purifier is used, especially if the water contains a large percentage of 



Q£> 






QG 



ft 



ILLH » BUILtM 

±± 

— ■ FEEO T MAIN — ■ 




INJECTOR MAIN 



— I-J [ FEED PUMP | | FECI 



I J I | INJECTOR—l X^I 1 

4—1 WATER 1*1 ' 



COLD WATER SUPPLY 



HEATER / I 



"•».»!»* X " 



tass*-- 



Fig. 371. Feed Water Piping; Now Condensing Plant. 



calcium sulphate. The feed is then subjected to boiler pressure and 
temperature and the greater part of the impurity precipitated before it 
enters the boiler. Closed heaters are often used in place of open 
heaters. When the supply is not under head a closed heater is usually 
preferred and is placed between the pump discharge and the feed 
main. 

In condensing plants the feed piping is similar to that in non- 
condensing plants, except that if exhaust steam is used for heating 



©0 o© 



ex 






n 




a 










Fig. 372. Feed Water Piping; Condensing Plant. 

purposes it is supplied by the auxiliaries, such as feed pumps, stoker 
engines, condenser engines, and other steam-using appliances. 

In plants having a number of boilers it is customary to run a feed 
main or header the full length of the boiler room and connect it to 



PIPING AND PIPE FITTINGS 



649 



each boiler by a branch pipe. This main may be a simple header or 
in duplicate or of the " loop " or " ring " type. Horizontal tubular 
boilers are frequently arranged in one battery with the feed main run 
along the fronts of the boilers just above the fire doors. Water-tube 
boilers are generally set in a battery, and as the arrangement above 
would block the passageway between the batteries, the main is run 
either above or under the settings, the former being the more common. 
Where a single header is used, the feed pumps are sometimes placed so 
as to feed into opposite ends of the main, which is then cut into 
sections by valves. Another arrangement is to place the pumps so as 
to feed into the middle of the header. With the loop arrangement the 
main is ordinarily cut into sections by valves so that the water may be 
sent either way from the pumps and any defective section cut out. 
With duplicate mains a common arrangement is to place one main 
along the front of the boiler and the other at the rear or both over- 
head as in Fig. 359. Sometimes one main is placed in the passageway 
below the boiler setting and the other on top. 

Standard wrought-iron pipe is usually used for pressures under 100 
pounds and extra heavy pipe for greater pressures. The pipes and 
fittings from boiler to main are frequently of brass, and preferably* so, 
since brass withstands corrosive action much better than iron or steel. 
Flanged joints should be used in all cases, since the pockets formed by 
the ordinary screwed joints hasten corrosion at those points. {Power, 
June, 1902, p. 4.) 

Fig. 373, A to E, illustrates the various combinations of check valve, 
stop valves, and regulating valve in steam boiler practice. The 



Fig. 373. Different Arrangements of Valves in Feed Water Branch Pipes. 



simplest arrangement and one sometimes used in plants operating 
intermittently is shown in A. Here there are but two valves between 
the boiler and the main, the check being nearest the boiler and the stop 
valve at the main. The stop valve performs both the function of 
cutting out the boiler and of regulating the water supply. This 



650 



STEAM POWER PLANT ENGINEERING 



arrangement is not recommended, as any sticking or excessive leaking 
of the check valve will necessitate shutting down the boiler. B shows 
the most common arrangement. Here the check valve is placed 
between the regulating valve and a stop valve as indicated. This 
permits a disabled check to be easily removed while pressure is on the 
boiler and the main. E shows an arrangement whereby both check 
and regulating valve may be removed, and is particularly adapted to 
boilers operating continuously where the regulating valve is subjected 
to severe usage. In this case the stop valves are run wide open and 
are subjected to no wear. The regulating valve most highly recom- 
mended is a self-packing brass globe valve with regrinding disk. The 
check valve is ordinarily of the swing check pattern with regrinding 
disk, Fig. 384 (C). Modern practice recommends an automatic water 
relief valve in the discharge pipe immediately adj acent to each pump to 
prevent excessive pressure in case a valve is accidentally closed in 
by-passing or in changing over. 

344. Flow of Water through Orifices, Nozzles, and Pipes. — Ber- 
noulli's theorem is the rational basis of most empirical formulas for 
the steady flow of a fluid from an up-stream position n to a down- 
stream position m, thus ("Mechanics of Engineering/ ' Church, p. 706): 



+ 



TV 



+ Z m = 



Pn 

y 



+ 



+ z t 



fall losses of head 1 
-[ occurring between >•> 



[n and 



m 



r — r "m — T - 

y 2g y 2g 

in which 

V = velocity in feet per second at the point considered. 
P = pressure in pounds per square foot. 
Z = potential head in feet of the fluid. 

y = density of the fluid, pounds per cubic foot. 

g = acceleration of gravity. 

V 2 



(156) 



Each loss of head will be of the form 



coefficient of resistance to be determined experimentally 
head due to skin friction is expressed : 

7 v 2 

in which 

/ = the coefficient of friction of the fluid in the pipe. 
I = length of the pipe in feet. 
d = diameter of the pipe in feet. 
Other notations as in (156). 



K— , in which K is the 
2g 9 

The loss of 



(157) 



PIPING AND PIPE FITTINGS 651 

Discharge from a circular vertical orifice with sharp corners: 

Q = CA V2gh } (158) 

in which 

Q = cubic feet per second. 

C = coefficient, varying from 0.59 to 0.65 (Merriman, Treatise on 

Hydraulics, p. 118). 
A = area of the orifice, square feet. 
h = head of water in feet. 
g = acceleration of gravity =32.2. 

Discharge from short cylindrical nozzles three diameters in length, with 
rounded entrance (" Mechanics of Engineering," Church, p. 690) ; 

Q = 0.815 A V2~gh. (159) 

Discharge from short nozzles with well-rounded corners and conical 
convergent tubes, angle of convergence 13J degrees (Church, p. 693): 

Q = 0M-A\/2gh. (160) 

Discharge from cylindrical pipe under 500 diameters in length 
(Church, p. 712): 

Q = 6 - 3 v/ ( i + £ + ? fl ' (161) 

in which 

/ = coefficient of friction. 

Other notations as above. 

/ varies with the nature of the inside surface, the diameter of the 
pipe, and the velocity of flow. 

Discharge through very long cylindrical pipes (" Mechanics of Engi- 
neering," Church, p. 715): 

Q = 3.15 y/^ • (162) 

Loss of head due to friction in water pipes. Weisbach's formula is 
as follows: 

ou 4+ 0J)lM^ (163) 

\/V J 5.367 d 



H=(0 



in which 

H = friction head in feet. 

V = velocity in feet per second. 

L = length of pipe in feet. 

d = diameter of pipe in inches. 



652 



STEAM POWER PLANT ENGINEERING 



TABLE OF THE COEFFICIENT / FOR FRICTION OF WATER IN CLEAN 

IRON PIPES. 

(Abridged from Fanning.) 



Velocity in 
Ft. per Sec. 


Diam. 
= | in. 


Diam. 
= 1 in. 


Diam. 
= 2 in. 


Diam. 
= 3 in. 


Diam. 
= 4 in. 


Diam. 
= 6 in. 


Diam. 

= 8 in. 


= .0417 ft. 


= .0834 ft. 


= .1667 ft. 


= .25 ft. 


= .333 ft. 


= .50 ft. 


= .667 ft. 


0.1 


.0150 


.0119 


.00870 


.00800 


.00763 


.00730 


. 00704 


0.3 


.0137 


.0113 


850 


784 


750 


720 


693 


0.6 


.0124 


.0104 


822 


767 


732 


702 


677 


1.0 


.0110 


.00950 


790 


743 


712 


684 


659 


1.5 


.00959 


.00868 


.00757 


.00720 


.00693 


.00662 


. 00640 


2.0 


.00862 


810 


731 


700 


678 


648 


624 


2.5 


795 


768 


710 


683 


662 


634 


611 


3.0 


.00753 


.00734 


.00692 


.00670 


.00650 


. 00623 


.00600 


4.0 


722 


702 


671 


651 


631 


607 


586 


6.0 


689 


670 


640 


622 


605 


582 


562 


8.0 


663 


646 


618 


600 


587 


562 


544 


12.0 


630 


614 


590 


582 


560 


540 


522 


16.0 


.00618 


.00600 


.00581 


.00570 


.00552 


.00530 


.00513 


20.0 


615 


598 


579 


566 


549 


525 


508 



Velocity in 
Ft. per Sec. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


Diam. 


= 10 in. 


= 12 in. 


= 16 in. 


= 20 in. 


= 30 in. 


= 40 in. 


= 60 in. 


= .833 ft. 


= 1.00 ft. 


= 1.333 ft. 


= 1.667 ft. 


= 2.50 ft. 


= 3.333 ft. 


= 5. ft. 


0.1 


.00684 
673 
659 


. 00669 
657 
642 


.00623 
614 
603 










0.3 


.00578 
567 








0.6 


.00504 


.00434 


.00357 


1.0 


643 


624 


588 


555 


492 


428 


353 


1.5 


.00625 


.00607 


.00572 


.00542 


.00482 


.00421 


.00349 


2.0 


609 


593 


559 


529 


470 


416 


346 


2.5 


596 


581 


548 


518 


460 


410 


342 


3.0 


.00584 


.00570 


.00538 


.00509 


.00452 


.00407 


. 00339 


4.0 


568 


553 


524 


498 


441 


400 


333 


6.0 


548 


534 


507 


482 


430 


391 


324 


8.0 


532 


520 


491 


470 


422 


384 


320 


12.0 


512 


500 


478 


457 


412 


377 


.00313 


16 


00502 


00491 


00470 


00450 


00406 


00370 




20 


498 


485 

























William Cox {American Machinist, Dec. 28, 1893) gives a simple 
formula which gives almost identical results: 



H = 



(4 V 2 + 5 7-2) L 

1200 d 
Notations as in (163). 



(164) 



Loss of head due to friction of fittings. Formulas (161) to (164) are 
based on the flow of water through clean straight cylindrical pipes. 



PIPING AND PIPE FITTINGS 653 

Where there are bends, valves, or fittings in the line the flow is 
decreased on account of the additional resistance. 

These frictional losses are conveniently expressed in feet of water, 
thus: 

H = C^P- (165) 

C having the following values: 

Angles. Class of Valve. 



45 degrees. 90 degrees. Gate. Globe. Angle, 

C 0.182 0.98 0.182 1.91 2.94 

Example: Determine the pressure necessary to deliver 200 gallons 
of water per minute through a 4-inch iron pipe line 400 feet long, fitted 
with four right-angle elbows and two globe valves. The water is to 
be discharged into an open tank. 

A flow of 200 gallons per minute gives a velocity of 

?= 5 feet per second (7.48 = number of gallons per 



7.48 X 60 X 12.72 

cubic foot, and 12.72 = internal area of the pipe, square inches) 

From the preceding table, / = 0.618 for 7 = 5. 

From (165), 

Resistance head of 4 elbows = 0.98 X -=^ X 4 = 1.52 feet. 

64.4 

Resistance head of 2 globe valves: 

1.91 X -^- X 2 = 1.48 feet. 
64.4 

Resistance head of all fittings: 

1.52+ 1.48 = 3 feet. 

Substitute 7 = 5, L= 400, and d = 4 in (164). 

(4 X 5 2 + 5 X 5 - 2" 



"-( 



400 



1200 X 4 
10.25 feet, resistance head of the pipe. 

Total resistance head = 10.25 + 3 = 13.25 feet of water, or 5.75 
pounds per square inch. 

Example: How many gallons of water will be discharged per minute 
through above line with initial pressure of 100 pounds per square inch, 
and what will be the pressure at the discharge end? 

Since / depends upon the unknown 7, we may put / = 0.006 for a 
first approximation and solve for 7; then take a new value of / and 
substitute again, and so on. 



654 



STEAM POWER PLANT ENGINEERING 



Substitute /= 0.006, d= -^- ,h = 100X2.3 = 230, and Z = 400 in (162) 



Q = 3.15 



Vo. 



33 5 + 230 



.006 X 400 

= 1.95 cubic feet per second, corresponding to a 
velocity of 22 feet per second. 

From the preceding table, 

/ = 0.00548 (by interpolation) for V = 22 feet per second. 

From (165) the friction of 4 elbows and 2 globe valves is found to 
be 58 feet for V = 22. 

From (164) a resistance head of 58 feet of water for V = 22 is 
found to be equivalent to 136 feet of straight pipe, thus: 
4 X 22 2 X 5 X 22-2 



58 



-■ 



1200 X 4 



.), 



L = 136. 
Substitute /= 0.0548, I = 400 + 136 = 536 in (162): 

Q=-3.15v/°- 335X230 
V 0.0058 X 536 

= 1.74 cubic feet per second, corresponding to a velocity 

of 19.3 feet per second. 

= 780 gallons per minute. 

If greater accuracy is necessary determine / and L for V = 19.3 
and proceed as above. 

The total friction head may be determined from (164) thus: 
4 X 19.3 2 + 5 X 19.3-2 



H 



- 



536 



1200 X 4 
= 177 feet of water. 
= 77 pounds per square inch. 

The pressure at the discharge end will be 

100 — 77 = 23 pounds per square inch. 

Average power plant practice gives the following maximum veloci- 
ties of flow in water pipes : 



Size of Pipe in 
Inches. 


Velocity, Feet per 
Minute. 


Size of Pipe in 
Inches. 


Velocity, Feet per 
Minute. 


Hoi 

J to lj 
1J to 3 


50 
100 
200 


3 to 6 
Over 6 


250 
300-400 



PIPING AND PIPE FITTINGS 



655 



345. Stop Valves. — The valves used to control and regulate the 
flow of fluids are the most important element in any piping system. 
A good valve should have sufficient weight of metal to prevent 
distortion under varying temperature and pressure, or under strains 
due to connection with the piping; the seats should be easily repaired 
or renewed; there should be no pockets or projections for the accu- 
mulation of dirt and scale, and the valve stem should permit of easy 
and efficient packing. Stop valves are made in such a variety of 
designs that a brief description will be given of only a few of the best- 
known types. 

Fig. 374 shows a section of an ordinary globe valve, so called because 
of the globular form of the casing. This type of valve is the most 





Fig. 374. A Typical Globe Valve, 
Screw-Top, inside Screw. 



Fig. 375. A Typical Globe Valve, 
Bolt-Top, outside Screw. 



common in use. Globe valves are designated as (1) inside screw and 
(2) outside screw, according as the screw portion of the stem is inside 
the casting, Fig. 374, or outside, Fig. 375. The top or bonnet may be 
screwed into the body of the valve, Fig. 374, or bolted, Fig. 375. The 
smaller sizes, three inches and under, are usually of the screw-top type 
and the larger of the bolt-top type. Valves with outside yoke and 
screw are preferable to the other in that they show at a glance whether 
the valve is open or closed, an advantage in changing from one section 
to another. The disks are made in a variety of forms, the material 



656 



STEAM POWER PLANT ENGINEERING 



depending upon the nature of the fluid to be controlled. Thus, for 
cold water, hard rubber composition gives good results; for hot water 
and low-pressure steam, Babbitt metal; for high-pressure steam, 
copper or bronze; and for highly superheated steam, nickel. The 
valve bodies are of brass for sizes under three inches, cast iron for the 
larger sizes and ordinary pressures and temperatures, and cast steel or 
semi-steel for high temperatures and pressures. Globe valves should 
always be set to close against the pressure, otherwise they could not 
be opened if the valves should become detached from the stem. Globe 
valves should never be placed in a horizontal steam return pipe with 





Fig. 376. A Typical Gate 

Valve, Solid-Wedge, Screw- 
Top, outside Screw. 



Fig. 377. A Typical Gate Valve, 
Solid- Wedge, Bolt-Top, inside 
Screw. 



the stem vertical, because the condensation will fill the pipe about 
half full before it can flow through the valve. Globe valves that are 
open all the time are preferably designed with a self-packing spindle, as 
in Fig. 375, in which the top of shoulder C can be drawn tightly 
against the under surface of bonnet S, thus preventing steam from 
leaking past the screw threads while the spindle is being packed. 

Figs. 376 to 379 show different types of gate or straight way valves. 
These valves offer little resistance to the flow of steam or liquid passing 
through them, and are generally used in the best class of work. 
Fig. 376 shows a section through a solid-wedge gate valve with outside 






PIPING AND PIPE FITTINGS 



657 



screw and yoke. This form of outside screw and yoke with stem 
protruding beyond the hand wheel is a perfect indicator to show 
whether the valve is open or shut, as the hand wheel is stationary 
and the spindle rises in direct proportion to the amount the valve is 
opened. For these reasons outside screw valves are preferable for 
high-pressure work and especially for the larger sizes. The seats are 
made solid, or removable, and of various materials for different pres- 
sures and temperatures. Fig. 378 shows a section through a split- 
wedge gate valve with parallel faces and seats. For the sake of illus= 





Fig. 378. A Typical Gate Valve, Split- 
Wedge, Bolt-Top, inside Screw. 



Fig. 379. 



Ludlow Angle Valve, Gate 
Pattern. 



tration this valve is fitted with inside screw. In this design the 
spindle remains stationary so far as any vertical movement is con- 
cerned, and the gate or plug being attached to it by means of a 
threaded nut rises into the bonnet when the spindle is revolved. 
It is impossible to tell by its appearance whether this form of valve 
is open or closed. Valves with inside screw are adapted to situa- 
tions where there is considerable dirt and grit, since the screw is 
inclosed and protected and excessive wear is thus avoided. Gate 
valves with split gates are more flexible than those with solid gates, 
and hence are less likely to leak. Fig. 379 shows the application 



658 



STEAM POWER PLANT ENGINEERING 



of the gate system to an angle valve. All high-pressure valves above 
8 inches in diameter should be provided with a small by-pass valve, 
as the pressure exerted against the disk or gate is very great when the 
valve is closed and the force required to move it is considerable. The 
by-pass valve also facilitates " warming up " the section to be cut in 
and is more readily operated than the main valve. 

346. Automatic Non-Return Valves. — Fig. 380 shows a section 
through an automatic non-return valve as applied to the nozzle of a 

steam boiler. As will be seen from the 
illustration it amounts to practically a 
large check valve with cushioned disk. 
The object of this device is the equali- 
zation of pressure between the different 
units of the battery, the valve remaining 
closed as long as the individual boiler 
pressure is lower than that of the header. 
In case a tube blows out the valve closes 
automatically, owing to the reduction of 
pressure and prevents the header steam 
from entering the boiler. It acts also 
as a safety stop to prevent steam being 
turned into a cold boiler while men are 
working inside, because it cannot be 
opened when there is pressure on the 
header side only. To be successful, 
such a valve should not open until the 
pressure in the boiler is equal to that in 
the header; it should not stick and become inoperative nor chatter 
and hammer while performing its work. Referring to Fig. 380, tail 
rod E insures alignment and hence prevents sticking; steam space C 
acts as a dashpot to prevent hammering of the valve as it rises, and 
steam space D acts as a cushion and prevents hammering at closing. 
Lip F is made to enter the opening in the seat and reduce wire draw- 
ing across the seat. Fig. 358 shows the installation of a number of 
non-return valves at the Yonkers power house of the. New York Central 
Railway Company. 

347. Emergency Valves. — In large power plants it is customary to 
protect the various divisions of the steam piping by emergency valves 
which may be closed by suitable means at any reasonable distance 
from the valve. The simplest form of emergency stop is a weighted 
" butterfly " valve, which is to all intents and purposes a weighted 
check as illustrated in Fig. 385 (F) . The weight when supported, say 




Fig. 380. 



Anderson Non-Return 

Valve. 



PIPING AND PIPE FITTINGS 



659 



by a cord and pulley, holds the valve open; when the cord is cut or 
released the weight drops and forces the valve shut. The cord may 
lead to any convenient and safe distance from the valve. In applying 
this system of control to steam engines the valve is placed in the 
steam pipe just above the throttle and the weight held up by a lever 
controlled by the main governor or preferably by a separate gov- 
ernor. Should the engine exceed a certain speed, as in case of accident 
to the regular governor, the lever supporting the weight is tripped 
by the emergency governor and the valve is closed automatically. 
For high pressures a rotating plug valve or cock is preferred to the 
butterfly type, since it is balanced in all positions. Gate and globe 
valves may be converted into emergency valves by 
3 having the stems mechanically operated by electric 
motors, hydraulic pistons, and the like. Fig. 381 
shows a section through a Crane hydraulically oper- 
ated emergency gate valve. 



A 






Fig. 381. Crane 
Emergenc y Valve , 
Hydraulic. 



Fig. 382. Anderson Triple- 
Duty Emergency Valve. 



Fig. 383. Pilot Valve for 
Anderson Triple-Duty 
Emergency Valve. 



Fig. 382 shows a partial section through an " Anderson triple- 
duty " emergency valve, and Fig. 383 a section through the pilot 
valve. A steam connection from the main line to the top of a copper 
diaphragm holds the pilot valve closed because of the large area 
above the diaphragm. A steam pipe connection from underneath 
the emergency piston of the triple-acting valve also leads to the pilot 



660 



STEAM POWER PLANT ENGINEERING 



valve. In case a break occurs in the main steam line or branches, 
the pressure is removed from the top of the pilot valve, causing it to 
open, thus exhausting the pressure from beneath the emergency piston 
in the triple-acting valve. The boiler pressure on top of the emer- 
gency piston causes the valve to close. Pilot valves may be located 
at any desirable places, thus affording control from different points. 

In the " Locke automatic engine stop system " the stop valve is 
operated by an electric motor which is controlled by contact points 
operated by a speed-limit device. (See Power, August, 1907, p. 471, 
for a detailed description.) 

348. Check Valves. — Fig. 384, A to D, illustrate the different 
types of check valves in most common use. A is a ball check, B a cup 




f."'- A'. 


'1 / \ b sv>ssv l 




V^jSgr C^&p^ZZZl 


Tby^rfp "~ "" ~&£$srfff7 



(A) 



(B) (C) 

Fig. 384. Types of Check Valves. 



(oj 



or disk check, C a swing check, and D a weighted check. Occasionally 
the valve body is fitted with a valve stem and handle for holding the 
disk against its seat, in which it is designated as a stop check. In A 
and B the valve seat is parallel to the direction of flow and the valve 
is held in place by its own weight and by the pressure of the fluid in 
case of reverse flow. In the swing check the seat is at an angle of 
about 45 degrees to the direction of flow. The latter construction is 
preferred as it offers less resistance to flow and there is less tendency 
for impurities to lodge on the valve seat. By extending the hinge of 
the swing through the body of the valve, a lever and weight may be 
attached as in D and the check will not open except at a pressure 
corresponding to the resistance of the weight. It thus acts as a relief 
valve and at the same time prevents a reversal of flow. Stop checks 
are usually inserted in boiler feed lines close to the boiler, and when 
locked, act as any ordinary stop valve and permit the piping to be dis- 
mantled or the regulating valve to be reground without lowering the 
pressure on the boiler. Since the wear on check valves is excessive 
and necessitates frequent regrinding they are often mounted with 
regrinding disks, Fig. 384 (C), which may be " ground " against the seat 
without removing the valve from the line. 



PIPING AND PIPE FITTINGS 



661 



349. Blow-off Cocks and Valves. — The requirements of a good 
blow-off valve are that it shall furnish a free passage for scale and 
sediment, that it shall close tightly so as not to leak, and that it shall 
open easily without sticking or cutting. On account of the rather 
severe service to which such valves are subjected, they are made 
very heavy, with renewable wearing parts. 

Fig. 385 gives a sectional view of a Crane ferrosteel valve. The 
bonnet is easily taken off and the disk removed to be refaced or 
replaced by a new one. The old disk is repaired by pouring in a hard 
Babbitt metal and facing it off flush. The seats are of brass and oval 
on top to prevent scale lodging between them and the disk, and are so 
made that they may be removed; but it has been found in practice 
that there is not much cutting of the seat, the damage usually being 
confined to the softer Babbitt metal which faces the disk. 




Fig. 385. Crane Ferrosteel Fig. 387. 
Blow-off Valve. 



A Typical Blow-off 
Cock. 



Fig. 386. Faber Blow-off 
Valve. 



Fig. 386 gives a sectional view of a Faber valve. When the disk, 
which makes a snug fit in the body of the valve, is in the position 
shown, the boiler discharge is practically shut off and any sediment 
lying on the seat is cleaned off by a jet of steam or water. 

Fig. 387 shows a section through a typical blow-off cock of the 
straightway taper plug pattern with self-locking cam. Plug cocks are 
often used instead of valves on the blow-off piping. 

Current practice recommends the use of two valves, or rather one 
valve and one cock, in the blow-off line of each boiler. In most of 



662 



STEAM POWER PLANT ENGINEERING 













S— I 


L- 


„_.# 








^L 


If! 


V- 






[ 








■C 


^ 


r 








P d 


< 

z 
s 
2 

X 

u 

z 

10 

1 




I * 1 




, 




"0 J 


1 f 


z 






S — i — i 


I 




o 

\ 

LI 


1" 


■v 


t^M 








LI 

AC 

< 
hi 


M 
_l 












K 


«^ 




— Cr4 — » 

» i/ 1 ■ '- 


1 ^ 


! K ! 


! 


■ —.0-.li » 

-91 . 




L 








LJ 


8 -Or*—} 

Hi 




' '* 






l* 


^-.o-.n 




pld 


-L ' 


1 









PIPING AND PIPE FITTINGS 



663 



the large stations a blow-off valve and a blow-off cock are installed as 
indicated in Fig. 388. The number and size of blow-off cocks are 
usually specified by city or state legislation. 

350. Safety Valves. — Fig. 389 shows a section through the simplest 
form of safety valve. The valve is held on its seat against the boiler 
pressure by a cast-iron weight as 
indicated. This type has the ad- 
vantage of great simplicity, and 
can be least affected by tampering, 
since it requires so much weight 
that any additional amount which 
would seriously overload it can be 
quickly detected. For high pres- 
sure and large sizes of boiler this 
class of valve is entirely too cum- 
bersome. 

Fig. 390 shows the general de- 
tails of the common lever safety valve. 

The valve is held against its seat by a loaded lever, thereby enabling 
the use of a much smaller weight than the " dead weight " type, since 
the resistance is multiplied by the ratio of the long arm of the lever to 




Fig. 389. 



" Dead-weight 



Safety Valve. 




Fig. 390. Common Lever Safetv Valve. 



the short one. The proper position of the weight is determined by 
simple proportion. 

Fig. 391 shows a section through a typical pop safety valve in which 
the boiler pressure is resisted by a spring. This type of valve has 
practically supplanted all other forms. The boiler pressure acting upon 
the under side of valve V is resisted by the tension in spring S. As 
soon as the boiler pressure exceeds the resistance of the spring the 
valve lifts from its seat and the steam escapes through opening 0. 
The static pressure of the steam plus the force of its reaction in being 
deflected from the surface A holds the valve open until the pressure in 



664 



STEAM POWER PLANT ENGINEERING 



the boiler drops about 5 pounds below that at which the valve is lifted. 
The additional area of valve exposed to pressure when the valve lifts 

causes it to open with a sudden motion 
which has given it its name, and it also 
closes suddenly when the pressure has 
fallen. These valves are arranged so 
that the spring tension may be varied 
without taking them apart, and pro- 
vision is made for lifting the seats by 
means of a lever. The seats are of 
solid nickel in the best designs, to 
minimize corrosion. 

The commercial rating of a safety 
valve is based upon the area exposed 
to pressure when the valve is closed. 

The number and size of safety valves 
for a given boiler are ordinarily spec- 
ified by city or state legislation. 
The logical method for determining the size of safety valves is to 
make the actual opening at discharge sufficient to take care of all 
steam generated at maximum load. Most rules, however, are empir- 
ical and based on the extent of grate surface, thus: 

According to the Boiler Inspection Department 
Philadelphia, 22.5 G 




BOILER CONNECTION 



Fig. 391. 



Consolidated Pop Safety 
Valve. 



of the city of 
(166) 



in which P + 8 -6 

A = area of combined safety valves, inches. 

G = Grate area, square feet. 

P = boiler pressure, pounds per square inch gauge. 

According to the rule of United States Supervising Inspectors of 

Steam Vessels,* q 

A = — , for lever safety valves. 



A = 



C 
A = — , f or pop safety valves. 
o 
Other notations as in (166). 
Hutton's rule is (" Steam-Boiler Construction/') p. 470 : 
4 G 

Vp ' 

All notations as above. 

* Superseded 1908 by the following: 

A as above A - 0.2074 £ 

W = weight of steam per hour, lbs. 
P = Absolute steam pressure. 
See Power, Mar. 9, 1908, p. 480; Mar. 16, 1909, p. 520. 



(167) 
(169) 

(169) 



PIPING AND PIPE FITTINGS 



665 



The Consolidated Safety Valve Company's circular gives the fol- 
lowing rated capacity of its nickel-seat pop safety valves: 



Sizes in Inches. 


1 


1* 


1* 


2 


H 


3 


3* 


4 


4* 


5 5$ 


6 


Boiler Horse-power — 
From 


8 
10 


10 
15 


20 
30 


35 
50 


60 

75 


75 
100 


100 
125 


125 150 
150 175 


175 200 
200 275 


275 


To 


300 











351. Back-Pressure and Atmospheric Relief Valves. — These valves 
are for the purpose of preventing excessive back pressure in exhaust 
pipes. In non-condensing plants such valves are designated as back- 
pressure valves and in condensing plants as atmospheric relief valves. In 
the former the valve is usually adjusted so that a pressure of one to 





Fig. 392. Foster Back-Pressure Valve. 



Fig. 393. 



INLET 

Davis Back-Pressure Valve. 



five pounds above the atmosphere is necessary to lift it from its seat; 
in the latter the valve lifts at about atmospheric pressure. They are 
practically identical in construction, differing only in minor details. 
A slight leakage in the back-pressure valve is of small consequence, 
but in an atmospheric relief valve it may seriously affect the degree 
of vacuum and throw unnecessary work upon the air pump, hence it 
is customary to "water-seal" the latter. Fig. 392 shows a section 
through a typical back-pressure valve. The valve proper consists of 
a single disk moving vertically. The valve stem is in the form of a 
piston or dashpot which prevents sudden closing or hammering. The 
pressure holding the valve against its seat is regulated by a spring. 
When the back pressure becomes greater than atmospheric plus that 
added by the spring, the valve raises from its seat and relieves it. 

Fig. 393 shows a section through a Davis back-pressure valve, in which 
the resisting pressure is varied by means of a lever and weight. 



666 



STEAM POWER PLANT ENGINEERING 



Fig. 363 shows the application of a back-pressure valve to a typical 
heating system. 

Fig. 394 shows a section through a typical atmospheric relief valve. 
Opening B is connected to the exhaust pipe and opening A leads to 
the atmosphere. Under normal conditions of operation atmospheric 
pressure holds valve V against its seat. Water in groove S " water- 
seals " the seat and prevents air from being drawn into the condenser. 
In case the pressure in pipe B becomes greater than atmospheric it 
lifts valve V from its seat and is relieved. Piston P acts as a dash- 
pot and prevents the valve from slamming. 

Fig. 395 shows a section through an atmospheric relief valve in 
which the weight of the valve is counterbalanced or even over- 
balanced by an adjustable weight and lever, thereby permitting the 
valve to open at or below atmospheric pressure, as may be desired. 





Fig. 394. Crane Atmospheric Relief Valve. Fig. 395. Acton Atmospheric Relief Valve. 



352. Reducing Valves. — It is often necessary to provide steam at 
different pressures in the same plant, as in the case of a combined 
power and heating plant. To effect this result the reduction in pres- 
sure is accomplished by passing the steam through a reducing valve, 
which is but an automatically operated throttle valve. There are 
many different forms, the operation of all being based upon the same 
general principles. 

In the Kieley valve, Fig. 396, the low-pressure steam acts upon the 
top of flexible diaphragm D, and the weighted lever L (which may be 
adjusted to give the desired reduction in pressure) acts upon the other 
side. The movement of the diaphragm causes the balanced valve V 



PIPING AND PIPE FITTINGS 



667 



at the upper end of the spindle to open or close, as may be necessary 
to maintain the desired lower pressure. Inertia weights T and C 
prevent chattering. 





Fig. 396. Kieley Reducing 
Valve. 



Fig. 397. Foster Pressure 
Regulator. 



Fig. 397 shows a section through a class G Foster pressure regulator 
or reducing valve. In operation, steam enters at A and passes 
through the main valve port H to the outlet B. Steam at initial 
pressure passes through port C to chamber P and thence to the top 
of piston T through port L, opening the main valve U. Steam at 
delivery pressure passes through E and raises the diaphragm V 
against the pressure of spring R, allowing spring W to close the aux- 
iliary valve X. The pressure in chamber J is then equalized by the 
reduced pressure in ports G and the under side of piston X, and thus 
allows spring Y to close the main valve, which is then held to its seat 
by the initial pressure. Any reduction in delivery pressure is trans- 
mitted to diaphragm V, and permits spring to open auxiliary valve X, 
thereby admitting steam to the top of piston T, as previously 
explained. The delivery pressure is adjusted by screw D; thus 
increasing the tension of spring R increases the discharge pressure 
and vice versa. The adjustment once made, the delivery pressure 



668 



STEAM POWER PLANT ENGINEERING 



will remain constant, regardless of any variable volume of discharge 
or of the initial pressure, so long as the latter is in excess of the 
delivery pressure. W, Fig. 366, shows the application of a reducing 
valve to an exhaust steam heating system. Live steam is led to the 
valve through pipe A. It will be noted that the pipe leading from 
the valve to the heating system is much larger than the high-pressure 
supply pipe on account of the increase in volume of the low-pressure 
steam. Reducing valves should always be by-passed to permit of 
repairs without shutting down the system. Care should be taken in 
not selecting too large a reducing valve, as the valve lift is very small 
and the larger the valve the less will be the lift for a given weight of 
flow and consequently the greater the wire drawing and erosion of 
the valve seat. 

353. Foot Valves. — Whenever a long column of water is to be 
moved in either suction or delivery pipe it is customary to place a 
check valve near the lower end of the column to prevent the water 
from backing up when the pump reverses or shuts down. The check 
valve placed at the end of the suction pipe is called a foot valve. 
Any check valve may be used as a foot valve, though practice limits 
the choice to the disk or flap type as illustrated in Fig. 398. To pre- 
vent rubbish from destroying the action, a strainer or screen is gener- 






Fig. 398. Types of Foot Valves. 

ally incorporated with the body of the valve. A, Fig. 398, illustrates a 
single-flap, B a multi-flap and C a disk valve composed of a nest of 
small rubber valves. The single-flap are usually made in sizes J to 
6 inches, the multi-flap 7 to 16 inches, and the disk valve in all com- 
mercial sizes from f to 36 inches. For large sizes, 16 to 36 inches, the 
multi-disk valve is given preference, since a number of the disks may 
be disabled without destroying its operation. 

The Use and Abuse of Globe Valves: Power & Engr., Jan., 1909, p. 10. 
Gate Valves in Steam Pipe Lines: Power & Engr., Feb. 16, 1909, p. 320. 
Types of Check Valves and their Operation: Power & Engr., July 6, 1909, p. 11. 



CHAPTER XVL 

LUBRICANTS AND LUBRICATION. 

354. General. — The losses due to the friction of the working parts 
of machinery include considerably more than the mere loss of power, 
namely, the depreciation resulting from wear of bearings, guides, 
and other rubbing surfaces, and the expense arising from accidents 
traceable to excessive friction. The power absorbed in overcoming 
friction varies with the type of plant and the character of machinery 
and is seldom less than 5 per cent and often greater than 30 per cent 
of the total power developed. In large central stations these losses 
approximate 8 per cent and in weaving and spinning mills will average 
as high as 25 per cent. (Trans. A.S.M.E., 6-465.) These figures refer 
to properly lubricated plants operating under normal conditions. The 
proper selection of lubricant is therefore a very important problem, 
since, besides the cost of the lubricant itself, the loss in power and in 
wear and tear to machinery is no small item. A change of lubricant 
may frequently result in marked increase in economy of operation. 
The lubricants most commonly met with in power plant practice are 
conveniently classified as oils, greases, and solids, and are of animal, 
mineral, or vegetable origin. 

Reference Books: Archbutt and Deeley, Lubrication and Lubricants; Redwood 
Lubricants; W. M. Davis, Friction and Lubrication; Gill, Oil Analysis; Robinson, 
Gas and Petroleum Engines; Thurston, Friction and Lost Work; Gill, Engine Room 
Chemistry. 

355. Vegetable Oils. — Except for certain special purposes and for 
compounding with mineral oils these possess lubricating properties of 
little practical value, since they decompose at comparatively low tem- 
peratures and have a tendency to become thick and gummy. The vege- 
table oils sometimes employed are linseed, cottonseed, rape, and castor. 

Vegetable Oils: Power, May, 1906, p. 300; Archbutt and Deeley, Lubrication 
and Lubricants, p. 232; W. M. Davis, Friction and Lubrication, p. 28; Gill, Oil 
Analysis. 

356. Animal Fats. — Many animal fats have greater lubricating 
power than pure mineral oils of corresponding viscosity but are objec- 
tionable on account of their unstable chemical composition. They 
decompose easily, especially in the presence of heat, and set free acids 

669 



670 



STEAM POWER PLANT ENGINEERING 



which attack metals. They are seldom used in the pure state and 
are usually compounded with mineral oils. The animal products 
used in this connection are tallow, neat's-foot oil, lard, sperm, wool 
grease, and fish oil, the first named being the most important. In 
cylinder lubrication, especially in the presence of moisture, the addi- 
tion of 2 to 5 per cent of acidless tallow seems to make the oil adhere 
better to the metal surfaces and increases the lubricating effect, while 
the proportion is so small that ill effects from corrosion or gumming 
are scarcely perceptible. 

Animal Fats: Archbutt and Deeley, Lubrication and Lubricants, p. 323; Gill, 
Oil Analysis, p. 44; Wright, Analysis of Oils, p. 193; Andes, Animal Fats. 

357. Mineral Oils. — These are all products of crude petroleum and 
form by far the greater part of all lubricants. They present a wider 
range of lubricating properties than those derived from animal or 
vegetable sources, the thinnest being more fluid than sperm and the 
thickest more viscous than fats and tallows. They are not easily 
oxidized, do not decompose, become rancid, or contain acids. 

Crude American petroleum of specific gravity 0.802 may yield 
the following commercial products. (" Gas and Petroleum Engines," 
W. Robinson.) 



Average 
Percentage. 



Specific 
Gravity. 



Boiling 

Point, 

Degrees 

F. 



Light Oils. 

C Cymogene 

Petroleum ether < Rhigolene 

r Gasoline 

( C. Naphtha 

Petroleum spirit < B. Naphtha 

( A. Naphtha (benz.) 

Burning oils, kerosene. jordtnaTy kerosene". '. '. '. '. 

Fuel oils {For making oil gas; fuel 

C Lubricating oils 

Heavy oils < Paraffin wax 

( Residium 



traces 

0.1 

1-1.5 

10 
2-2.5 
2-2.5 

12-20 
40-55 



17.5 

2 
5-10 



0.590 
625-. 631 
635-. 658 

680-.700 

717-.72 
742-. 745 

780-.785 
800-.810 



0.85 

885-. 920 
908 at 60 
deg. F. 



32 
64 

85-155 

140-212 
175-245 
212-265 

300-570 
300-680 
and up- 
wards 



480 and 
upwards 



Mineral lubrication oils may be classified as 

(1) Distilled oils, which are produced by distillation from crude 
petroleum and made pale, amber colored, and transparent by treat- 
ment with acid and alkali. 



LUBRICANTS AND LUBRICATION 671 

(2) Natural oils, which are prepared from crude petroleum, from 
which grit, suspended and tarry impurities have been removed. They 
are dark and opaque and are rich in lubricating properties. 

(3) Reduced oils, or heavy natural oils, from which the lighter hydro- 
carbons have been evaporated and from which the tarry residue has 
been removed by nitration. 

Mineral Lubricants: Engr. U.S., July 1, 1904, pp. 466, Vol. 44 (1907), pp. 241, 369, 
542, 585; National Engr., Jan., 1905, p. 19; Eng. Mag., June, 1904, p. 455; Power, 
March, 1906, p. 146. 

358. Solid Lubricants. — Dry graphite, soapstone, and mica are 
sometimes used as lubricants, though they are usually mixed with 
grease or oils. They cannot easily be squeezed or scraped from 
between the surfaces, and are consequently suitable where very great 
weights have to be carried on small areas and when the speed of rub- 
bing is not high. The coefficient of friction of such lubricants is high, 
and when economy of power is essential better results may be secured 
by the use of liberally proportioned rubbing surfaces and liquid lubri- 
cants. Under certain conditions of pressure and speed these lubri- 
cants will sustain, without injury to the surfaces, pressures under which 
no liquid would work. 

Graphite: Trans. A.S.M.E., 13-374; Engng., Aug. 16, 1907; Sci. Am., May 11, 
1907; National Engr., Jan., 1904; Am. Mach., Dec, 1907, pp. 784, 934; Horseless 
Age, Jan., 1904, June 11, 1902, p. 712; Power, Dec., 1906, p. 758. 

359. Greases. — Under this name may be included the various - 
compounds which consist of oils and fats thickened with sufficient 
soap to form, at ordinary temperatures, a more or less solid grease. 
Those usually employed are lime, soda, or lead soaps, made with 
various fats and oils. " Engine " greases are thickened with a soap 
made from tallow or lard oil and caustic soda, and often contain 
neat's-foot oil, beeswax, and the like. For exceptionally heavy pres- 
sures, graphite, soapstone, and mica are sometimes added to the 
grease. 

Greases: Jour. Eng. Soc. West. Penn., March, 1904, p. 112; Railroad Gazette, 
July 8, 1904, p. 131; St. Ry. Jour., July, 1905, p. 95; see also text-books given in 
references at beginning of chapter. 

360. Qualifications of Good Lubricants. — A good lubricant should 
possess the following qualities: 

(1) Sufficient "body" to prevent the surfaces from coming into 
contact under conditions of maximum pressure. 

(2) Capacity for absorbing and carrying away heat. 



672 STEAM POWER PLANT ENGINEERING 

(3) Low coefficient of friction. 

(4) Maximum fluidity consistent with the " body " required. 

(5) Freedom from any tendency to oxidize or gum. 

(6) A high " flash point " or temperature of vaporization and a low 
congealing or " freezing point." 

(7) Freedom from corrosive acids of either metallic or animal 
origin. 

Lubricating oils are identified by certain tests which «,re used by 
refiners in grading and classifying the oils and by consumers in buy- 
ing them. These tests usually cover the following: 

(1) Identification of the oil, whether a simple mineral, animal or 
vegetable oil or a mixture. 

(2) Density or gravity. 

(3) Viscosity. 

(4) Flash point. 

(5) Burning point, fire test. 

(6) Acidity. 

(7) Coefficient of friction. 

(8) Cold test. 

361. Identification of Oil. — The chemical analysis of oils lies in 
the province of the chemist, but some of the characteristics may be 
readily determined by a few simple tests. To detect admixtures of 
fatty oils in mineral oil a small quantity is heated in a test tube for 
15 minutes with small pieces of either metallic sodium or caustic 
potash. If fatty oil is present, saponification takes place and the 
soap formed will rise to the top as a semi-solid mass and the amount 
may be estimated. Tarry matter may be detected by dissolving a 
small quantity of oil in from 10 to 20 times its bulk of gasoline; the 
tar and other insoluble matter will separate and collect at the bottom. 

Oil Testing and Specifications: Power, May, 1904, p. 302, Vol. 24 (1904), pp. 139, 
240, 302, 526, Vol. 26 (1906), pp. 145, 222, 300, 331, 407; Am. Mach., April 11, 1907, 
p. 525; Engr. U.S., Oct. 15, 1904, p. 724, Oct. 2, 1905, p. 657; Marine Engng., June, 
1903, p. 303; Chem. Engr. Nov., 1905, p. 10, Dec, 1905, p. 87, Jan., 1906, p. 141; 
Am. Gas Light Jour., Jan. 23, 1905; U.S. Cons. Repts., June, 1905; Sci. Am. Sup., 
Jan. 14, 1905. 

362. Gravity. — The density or specific gravity is conveniently 
determined by means of a hydrometer, which, in the oil trade, is 
graduated according to the Baume scale. The relationship between 
specific gravity and degrees Baume at a temperature of 60 degrees F. 
may be expressed: 

Specific gravity = 



130 + degrees Baume 






LUBRICANTS AND LUBRICATION 



673 



Table 88 gives the specific gravity and gravity Baume of a number 
of lubricating oils. 

Gravity: Power, March, 1904, p. 139; Robinson, Gas and Petroleum Engines, 
p. 474; Archbutt and Deeley, Lubrication and Lubricants, pp. 172-185; W. M. Davis, 
Friction and Lubrication, p. 34. 

TABLE 88. 

SPECIFIC GRAVITY AND GRAVITY BAUME OF A NUMBER OF LUBRICANTS. 



Water 

Cylinder oil 

Cylinder oil 

Heavy engine oil. . 
Medium engine oil 
Light engine oil. . . 
Castor machine oil 

Lard oil 

Sperm oil 

Tallow oil 

Cottonseed oil.. . . 

Linseed oil 

Castor oil (pure) . . 

Palm oil 

Rape-seed oil 

Spindle oil 



Specific Grav- 
ity. 



1.000 
.9090 
.8974 
.9032 
.9090 
.8917 
.8919 
.9175 
.8815 
.9080 
.9210 
.9299 
.9639 
.9046 
.9155 
.8588 



Gravity 
Baume\ 



10 

24. 

26 

25. 

24 

27 

27 

23 

29 

24. 

22 

19 

15 

25 

23 

33 



Flash Test, 
Degrees F. 



575 
540 
411 
382 
342 
324 
505 
478 
540 
518 
505 



405 
312 



363. Viscosity. — Viscosity may be denned as the degree of fluidity 
or internal friction of an oil. It is sometimes called the " body." It 
is determined by a viscosimeter. There are a number of different 
instruments for this purpose but no recognized standard instrument or 
method, so that " viscosity " conveys no meaning unless the name of 
the instrument, the temperature, and the amount of oil tested are 
given. Nearly all instruments are of the orifice type; that is, the 
viscosity of an oil is taken as the number of seconds required for a 
given amount, usually 50 cubic centimeters, to flow through an 
orifice at a given temperature. By " specific viscosity " is meant the 
ratio of the time required for the oil to run out to that of an equal 
quantity of water at 60 degrees F. The viscosity of engine oils is 
usually taken at 70 degrees F. and of cylinder oils at 212 degrees F. 

Viscosity: Trans. A.S.M.E., 9-369; Engr., Lond., Sept. 7, 1906, p. 344, June 12, 
1900, p. 633; Eng. Mag., June, 1907, p. 455; Machinery, May, 1903, p. 484; Power, 
May, 1904, p. 303, May, 1907, p. 293, March, 1906, p. 146. 

364. Flash Point. — The flash point is determined by heating a 
sample of oil in an open or closed cup at the rate of 15 degrees F. per 
minute until a spark will ignite the vapor. The temperature at which 






674 STEAM POWER PLANT ENGINEERING 

this occurs is the flash point. So much depends upon the extent of oil 
surface exposed, size of spark, distance spark is held from the oil at the 
time of ignition, and the dimensions of the cup, that there may be con- 
siderable variation in the flash point as obtained by different experi- 
menters. 

Flash Test: Power, April, 1906, p. 222; Robinson, Gas and Petroleum Engines, 
pp. 482-488; Archbutt and Deeley, Lubrication and Lubricants, pp. 187-191; W. M. 
Davis, Friction and Lubrication, p. 34; Gill, Oil Analysis, p. 36. 

365. Burning Point, or Fire Test. — By continuing the application 
of heat and noting the temperature at which the oil takes, fire and 
continues to burn, the burning point is obtained. The higher the 
temperature under which the oil must work the higher the fire test 
required, so that it will not decompose or volatilize. Too high a fire 
test gives an oil that does not atomize readily enough to reach all parts 
of the cylinder. 

Consult references under " Flash Test." 

366. Acidify. — The presence of free acid is determined by shaking 
up equal quantities of oil and water and testing with litmus paper. 
Another simple test is as follows: A small quantity of oil is placed in 
a test tube with a little cupric oxide (Cu 2 0) and subjected to a gentle 
heat for three or four hours. The reaction with the copper turns the 
solution green if fatty acid is present and blue if vegetable acid is 
present. 

Acidity: Power, April, 1906, p. 222; Archbutt and Deeley, Lubrication and 
Lubricants, pp. 215-218; Gill, Oil Analysis, p. 74. 

367. Cold Test. — The " cold test " is the temperature at which the 
oil will just flow. The sample is solidified by means of a freezing 
mixture and the temperature noted when it softens sufficiently to flow. 

Cold Test: Robinson, Gas and Petroleum Engines, p. 481; Archbutt and Deeley, 
Lubrication and Lubricants, pp. 195, 200-6; W. M. Davis, Friction and Lubrication, 
p. 28; Gill, Oil Analysis, p. 34; Redwood, Lubricants, p. 3; Power, March, 1906, p. 146. 

368. Friction Test. — The coefficient of friction as determined 
from friction-testing machines is useful in obtaining a comparison of 
oils under the test conditions, but gives little information concerning 
the action of the oil under the widely different conditions found in 
actual practice. Table 89 gives the physical properties of a number 
of lubricating oils, with their particular zone of application. 

Friction and Lubrication: Trans. A.S.M.E., 1-74, 6-136; Am. Mach., July 21, 
1904, p. 956, Jan. 23, 1902, p. 113, Sept. 10, 1903, p. 1303; Am. Elecn., Nov., 1905, 
p. 557; Engr., Lond., June 19, 1903, p. 631; Power, Dec, 1905, p. 748; National 
Engineer, Jan., 1905, p. 19; Mech. Engr., Nov. 30, 1907; Pro. Inst. Civ. Engr., 1901, 
p. 146; Machinery, Aug., 1903, p. 631; Proc. A.S.M.E., Nov., 1909, p. 1099. 






LUBRICANTS AND LUBRICATION 675 

369. Atmospheric Surface Lubrication. — In a general sense all 
journals, slides, and "atmospheric" surfaces should be lubricated with 
straight mineral oils (as free from paraffin as possible), except when 
in contact with considerable water, in which case it is advisable to add 
20 to 30 per cent of lard oil. Vegetable oils, paraffin oils, and animal 
oils (except lard oil as above stated) are not recommended for general 
engine and dynamo service. The test requirements of a number of 
classes of lubricants are outlined in Table 89 and represent current 
practice. Bearings, guides, and all external rubbing surfaces may be 
lubricated in a number of ways. (1) They may be given an inter- 
mittent application of oil, as, for example, with an oil can; (2) they may 
be equipped with oil cups with restricted rates of feed; and (3) they 
may be flooded with oil. The relative lubricating values of the systems 
have been estimated approximately as follows {Power, December, 1905, 
p. 750) : 



Intermittent. . . . 
Restricted feed. 
Flooded bearing. 



Coefficient of Fric- 
tion. 



Comparative 
Value. 



0.01 and greater 72 and less 

0.01 to 0.012 79 to 86 

0.00109 100 



370. Intermittent Feed. — Intermittent applications are ordinarily 
limited to small journals, pins, and guides which are subject to light 
pressures and which do not easily permit of oil or grease cups, as, for 
example, parts of the valve gear of a Corliss engine, governors and 
link work. On account of the labor attached and the frequent doubt 
about the oil reaching the wearing surfaces this method of lubrication 
is limited as much as possible even in the smallest plants. 

371. Restricted Feed. — In the- average power plant the major part 
of the lubrication is effected by means of oil cups which are filled at 
intervals by hand or by mechanical means, the oil being fed from the 
cup by drops, according to the requirements. 

372. Oil Bath. — In large power plants the principal journals and 
wearing parts are supplied with a continuous flow of oil which com- 
pletely " floods " the rubbing surfaces. The oil is forced to the various 
parts either by gravity from an elevated tank or by pressure from a 
pump. After the oil leaves the bearings it flows into collecting pans, 
thence into a receiving and filtering tank, and finally is pumped back 
into an elevated reservoir and used over and over again. The little 
lost by leakage and depreciation is replenished by the addition of new 
oil to the system. 



676 



STEM! POWER PLANT ENGINEERING 



TABLE 89. 
PHYSICAL CHARACTERISTICS OF A NUMBER OF LUBRICANTS. 

(Power, December, 1905, p. 750.) 



Kind of Oil. 


Use and Adaptation. 


O 


8 Q 




to $ 


Viscosity 
at 70 De- 
grees. 


High-pressure cylinder 
oil. 


For steam cylinders using dry 
steam at pressures from 110 
to 210 pounds. 


25 

to 

24.5 


30 


600 
to 
610 


645 

to 

660 


175 

to 

205 


General cylinder oil . . 


For steam cylinders using dry 
steam at 75 to 100 pounds. 
For air compressor cylinders 
when made from steam-re- 
fined mineral stock and when 
viscosity is 200. 


26 

to 

25.5 


30 


550 
to 

585 


600 
to 
630 


180 
to 
190 


Wet cylinder oil. 
(Remark 1.) 


For use where the steam is moist, 
especially in compound and 
triple expansion engines. 


25.8 

to 
25.3 


30 


560 
to 
585 


600 
to 
630 


150 
to 

185 


Gas engine cylinder oil. 
(Remark 2.) 


For gas engine cylinders. Neu- 
tral mineral oil compounded 
with an insoluble soap to give 
body. 


26.5 


30 


320 


350 


300 


Automobile gas engine 
oil. (Remark 3.) 


For automobile gas engines and 
similar work. 


29.5 


30 


430 


485 


195 


Heavy engine and 
machinery oils. 


For heavy slides and bearings, 
shafting, and horizontal sur- 
faces. 


30.5 

to 
29.5 


30 


400 


440 
to 
450 


170 

to 
195 


General engine and 
machine oils. 


For high-speed dynamos and 
machines. 


30.8 
to 
30 


30 


400 
to 
420 


450 
to 
470 


175 

to 
190 


Fine and light machine 
oils. 


For fine work, from printing 
presses to sewing machines 
and typewriter oils. With a 
cold test of 25° to 28° and a 
viscosity of 140° this makes 
an excellent spindle oil. 


32.5 

to 
30.2 


30 


400 


440 


110 
to 
160 


Cutting and heat dis- 
sipating oils. 
(Remark 4.) 


For cutting tools, screw cutting 
and similar work. 


27 
to 
23 


30 


410 
to 
420 


475 
to 

480 


210 
to 
175 




For ice machinery 


30.2 





200 


225 


165 








Wet service and marine 
oils. (Remark 4.) 


For marine service, or where a 
great deal of moisture must 
be handled. 


28 


30 


430 


475 


230 




They are used in special work 
and for heavy pressures mov- 
ing at slow velocities. 




















Remark 1. — May contain not over 2 to 6 per cent of refined acidless tallow oil in the high- 
pressure oils and not over 6 to 12 per cent in the low-pressure oils. 

Remark 2. — The reason for using an insoluble soap such as oleate of aluminum is that it 
is impossible to decompose the soap with a high heat ; the soap, although not a lubricant, is a 
vehicle for carrying some oil. 

Remark 3. — Owing to a lack of body, this oil will not interfere with the sparking by depos- 
iting carbon on the platinum point. 

Remark 4. — May contain 30 to 40 per cent of pure strained lard oil. 



LUBRICANTS AND LUBRICATION 



677 



373. Oil Cups. — Fig. 399 illustrates the application of sight-feed oil 
cups to the crosshead and slides of a reciprocating engine. The oil is 

fed into the cups by hand and 
gravitates to the rubbing surfaces, 
the rate of flow being regulated by 
a needle valve. Cups A and B 
feed directly to the crosshead 
guides, but the oil from cup D 
flows to the bottom orifice 0, 
from which it is wiped by a metal- 
lic wick S and carried by gravity 
to the wrist pin. 

374. Telescopic Oiler.— Pig. 400 
shows the application of a tele- 
scopic oiler to a crosshead and 
guides. and C are sight-feed 
oil cups, the former feeding directly 
to the top guide through the tube 
S. The oil from C flows by 
gravity through the swing joint into the telescopic tubes P, R and 
thence to the pin through the lower swing joint as indicated. As 




Fig. 399. 



Oil Cup Lubrication, Hand 
Filled. 





Fig. 400. Nugent 's Telescopic Oiler. 

the crosshead moves back and forth, the pipe P slides into and out of 
pipe R, the oil being thus conducted directly to the pin without wasting. 



678 



STEAM POWER PLANT ENGINEERING 



A device of this type installed on a high-speed automatic engine at the 
Armour Institute of Technology has been in operation for three years 
without cost for repair or renewal. 

375. Ring Oiler. — Small high-speed engines are often oiled by the 
oil-ring system ; as illustrated in Fig. 401. The shaft is encircled by 




M/MMUWBZZ* 



rj 4,^ 



<j 



Fig. 401. Oil Ring Lubrication. 



O O Ol 



OOO 



3 



Fig. 402. Centrifugal Oiler. 



several loose rings which dip into a bath of oil in the base of the 
pedestal or frame and, rolling on the shaft as it turns, carry oil to 
the top of the shaft where it spreads to the bearings. In some cases 
the rings are replaced by loops of chain. 

Ring Lubrication: Zeit. d. Ver. Deutscher Ing., Aug. 10, 1907. 




Fig. 403. Pendulum Oiler. 



376. Centrifugal Oiler. — Fig. 402 illustrates the application of a 
centrifugal oiler to a side-crank engine. The oil supply is regulated by 



LUBRICANTS AND LUBRICATION 



679 



the sight-feed cup C and flows by gravity to the pipe P in line with 
the center of the crank shaft. Centrifugal force throws the oil out- 
ward through pipe B to the center of the pin D, which is drilled longi- 
tudinally and radially so as to distribute the oil upon the bearing 
surface. 

377. Pendulum Oiler. — Fig. 403 illustrates the application of a 
pendulum oiler to the crank pin of a center-crank engine. Oil cups 
and pendulum P are fastened to the crank shaft S by trunnion T. 
The pendulum holds the cup vertical, since the friction of the trunnion 
is not sufficient to revolve it. Oil flows along the center of the crank 
shaft under the head of oil in cups and is thrown outward to bear- 
ing B by centrifugal force. 



OVER FLOW 




FLEXIBLE fl ^O 
FILLING 
PIPE 



r 



BASEMENT FLOOR LINE 



Fig. 404. Simple Gravity Feed System. 



378. ** Splash " Oiling. — In some high-speed engines the crank, con- 
necting rod, and crossheads are inclosed by a casing, the bottom of 



680 STEAM POWER PLANT ENGINEERING 

which is filled with oil to such a depth that at each revolution of the 
crank, the end of the connecting rod is partly submerged. The result 
is that the oil is splashed into every part of the chamber, and the 
crank pin, crosshead pin, and erosshead slides practically run in an oil 
bath. 

379. Gravity Oil Feed. — Fig. 404 illustrates a simple gravity oil feed 
system. The oil to the engine is supplied from the oil tank by pipe D 
under pressure corresponding to the height of the tank above the oil 
cups. After performing its function the oil gravitates to the filter and 
from the latter to the oil reservoir, from which it is pumped back to 
the supply tank, the overflow being returned to the reservoir through 
pipe N. Operation is interrupted only when new oil is to be added to 
the system from the barrel through the flexible filling pipe. In case 
the oil tank is put out of commission, or the supply pipe becomes 
clogged, full pump pressure may be used by closing valves R and S 
and opening valve E. The make-up oil is small in amount compared 
to the quantity circulated. The reclaiming and purifying of the oil 
are essential if the bearings are to be flooded, otherwise the cost of oil 
would be prohibitive. At the power house of the South Side Elevated 
Railway the daily circulation (24 hours) of engine oil is approxi- 
mately 1500 gallons. The make-up oil amounts to eight gallons. 

An objection sometimes made to the above system is that the vary- 
ing heights of oil in the supply tank may cause considerable variation 
in pressure at the oil cups, causing them to feed faster when the tank 
is full and slower when the tank is nearly empty. This applies only 
to installations where the supply tank is filled intermittently. 

380. Low-Pressure Gravity Feed. — Fig. 405 shows the application 
of a low-pressure oiling system in which the level in the sight feeds is 
kept constant. A is the main supply tank, B l and B 2 the upper and 
lower gauges indicating the oil level, C the supply pipe running to the 
engines, and D a small standpipe closed at one end and vented near 
the top. The reservoir is supplied with oil by the valve marked " inlet." 
When the tank is filled the oil rises in the standpipe D a correspond- 
ing height. The inlet valve is then closed and the oil in the standpipe 
feeds down to the level of the sight feeds or to a point where the air 
will enter the bottom of the tank. This will be the constant oil level, 
since oil flows from the tank only in proportion to the amount of air 
admitted. A head of 6 inches has been found to give the best results. 
(Engineer, U. S., March 16, 1903, p. 243.) 

381. Compressed-Air Feed. — Fig. 406 shows diagrammatically the 
arrangement of the oiling system at the First National Bank Building, 
Chicago. The storage tank containing the supply of engine oil is 



LUBRICANTS AND LUBRICATION 



681 




Fig. 405. Low-Pressure Gravity Feed, Constant Head. 




SETTLING TANK 



I J 



klj 



OIL STORAGE 



H M - a _ - 



Fig. 406. Oiling System at the Power Plant of the First National Bank Building, Chicago 



682 STEAM POWER PLANT ENGINEERING 

under air pressure at all times except during the short periods when it 
is being filled with oil from the filter. The air pressure on the surface 
of the oil forces it to a manifold on the engine from which it is dis- 
tributed to the various oil cups. The oil flows from the different 
bearings to the returns tank located at the base of the engines. When 
the tank is filled air pressure is admitted and the oil forced to the 
settling tank, which has a capacity of about 400 gallons and is located 
near the ceiling. The oil is allowed to settle and the entrained water 
and foreign material are drained to waste. The oil gravitates from this 
tank to a series of Turner oil filters. When a new supply of oil is 
needed, valves A and B are closed and vent valve C opened, cutting 
off the supply of air and reducing the pressure to atmospheric. Valve D 
is then opened and oil flows from the filters to the storage tank. 

Lubricating Systems. — Lubrication of Line and Counter Shafting: Trans. A.S.M.E., 
10-810. Gravity Oil Systems: Power, Nov., 1902, p. 23, July, 1906, p. 409, June, 
1903, p. 305. Oiling Systems for Electric Engines: Elec. World, July 7, 1906, p. 26. 
Oiling System for Power Plants: Engr. U.S., March 16, 1903, p. 243, April 15, 1904, 
p. 278; National Engr., Feb., 1905, p. 10, March, 1905, p. 16. 

382. Cylinder Lubrication. — The test requirements for cylinder oils 
are outlined in Table 89, from which it will be seen that pure mineral 
oil fulfills practically all requirements for dry steam. In connection 
with moist steam, as in the low-pressure cylinders of compound engines, 
an addition of from 2 to 5 per cent of acidless tallow oil is recom- 
mended. Vegetable oils, beeswax, lard oil, degras (wool grease), and 
the like should never be used in compounding cylinder oils. The best 
cylinder oils are made from Pennsylvania stock. 

Cylinder oils must be forced to the parts requiring lubrication 
against the prevailing steam pressure, which is ordinarily accomplished 
by (1) cylinder cups, (2) hydrostatic lubricators, or (3) hand or power 
driven force pumps. 

383. Cylinder Cups. — A cylinder oil cup consists essentially of a 
steam-tight brass vessel fitted at the bottom with a pipe connection 
and valve. A screwed cap offers a means of introducing the lubri- 
cant into the cup. After the cap is in place the valve is opened and 
the cup is subjected to full steam pressure. The pressure in the cup 
being equal to that in the steam chest or cylinder, permits the lubri- 
cant to gravitate through the valve into the cylinder. 

Fig. 407 shows a section through an improved form of oil cup in 
which the oil feeds from the top instead of the bottom as is the case 
with the common form of cylinder cup. The vessel is attached to the 
steam chest or to the supply pipe below the throttle valve. Steam is 
admitted through opening B and, condensing, settles through the oil 



LUBRICANTS AND LUBRICATION 



683 



to the bottom. This raises the level of the oil until it begins to over- 
flow down the same passage by which the steam enters. This action 
is intensified by the fluctuation in steam pres- 
sure. The rate of feeding is regulated by valve 
C and tested by unscrewing plug F. If oil 
appears through opening G, the cup is feeding 
oil; if steam or water is emitted, the cup is 
empty. The cup is filled by means of plug E 
and the water drained at D. 

384. Hydrostatic Lubricators. — The most 
common method of cylinder lubrication is by 
means of hydrostatic lubricators of the sight-feed 
class, Fig. 408. The principle of operation is 
as follows: The lubricator is filled with cylinder 
oil by removing cap K, the height of oil ap- 
pearing in glass L. If water is present the 
oil floats on top as indicated. After the cap 




Fig. 407. Leyland Auto- 
matic Cylinder Cup. 



r^Oi 



K? 



is screwed in place the valves in the con- 
denser pipe are opened, subjecting the 
oil in the vessel to steam-pipe pressure. 
Steam is condensed in pipe C, filling tube 
B and part of C, thus adding to the steam 
pressure the pressure due to the weight 
of the water column. Valve F, which 
communicates with the top of the vessel 
by means of tube A, is opened wide, as 
is also the regulating valve I. The pres- 
sure at B being greater than that at A 
by an amount equivalent to the height of 
the water column, forces the oil through A 
and the " sight feed " S to the steam pipe. 
The rate of flow is controlled by the regu- 
lating valve /. As the oil flows from the 
vessel its space is occupied by condensed 
steam, the height of oil and water being 
visible in glass L. Owing to the small 
capacity of the lubricator it must be re- 
filled frequently. To reduce the amount 
of labor required with the above appa- 
ratus, independent sight feeds, Fig. 409, 
are sometimes used in connection with a 
central reservoir. Such an installation is shown diagrammatically in 




1X3 

Fig. 408. Common Hydrostatic 
Lubricator. 



684 



STEAM POWER PLANT ENGINEERING 



ffl 



^ 



Fig. 410. A condenser pipe leading from the steam main enters the 
bottom of the reservoir and the condensed steam fills up the reservoir 

as fast as the oil is fed out. The prin- 
ciple is the same as that of the simple 
hydrostatic lubricator. Oil is frequently 
injected by mechanical means under a 
steady pressure generated and governed 
independently of the steam. Two sys- 
n-.. f±± , - 1 terns are in common use, direct mechan- 

U == LP- gl |^ — | rMJl— TpT ical pump pressure and air pressure. 

385. Forced-Feed Cylinder Lubrica- 
tion. — Fig. 411 illustrates the " Roches- 
ter" simple feed automatic lubricating 
pump, which takes the oil by gravity 
from the reservoir through a sight-feed 
glass and forces it through a small pipe 
to the steam supply pipe. The pump 
entirely obviates the trouble due to in- 
termittent feeding and, being directly 
driven from the engine, runs at con- 
stant speed. The feed is uniform and 
independent of the pressure pumped against. The rate is deter- 
mined by the length of stroke of the pump piston, which is easily 
adjusted. 




^Oi 



Fig. 409. 



Lunkenheimer Sight-Feed 
Lubricator. 




Fig. 410. Central Hydrostatic Lubricator. 



With large engines multi-feed pumps are sometimes used, which 
force oil to the various valves as well as to the steam pipe. Fig. 412 



LUBRICANTS AND LUBRICATION 



685 



shows an arrangement of storage tank in connection with pump reser- 
voir to avoid the trouble of hand filling. 




Fig. 411. Rochester Forced-Feed Lubricator. 



H.P.Steam Pipe 



L.P. Steam Pipe To Rod 
To Rod >. , / 




Fig. 412. Forced-Feed Cylinder Lubrication. 



386. Siegrist System. — Fig. 413 shows an application of the 
Siegrist system of cylinder and engine lubrication. There are two 
storage tanks on the engine-room floor, one for cylinder oil and the 
other for engine oil, the distributing arrangements being the same in 
each case. The oil is pumped from each tank into a main pipe 
extending the length of the engine room and provided with branches 
at each point requiring lubrication. The oil pumps are actuated by 
steam and are of the duplex direct-acting type, provided with auto- 
matic governors which regulate the speed to suit the demand for oil. 



686 



STEAM POWER PLANT ENGINEERING 



5-j 

05 



!T-^-r^i.-Jr-_t: 



ii-.l--J 



I 
i 

— — i 

_] 1 i 



V2H— i-. 



j ^b ■ h - r 'Fol 





LUBRICANTS AND LUBRICATION 



687 




TO STEAM 



FEED REGULATOR 



Fig. 414. Siegrist Sight- Feed Lubricator. 



The cylinder oil is forced through a special sight-feed lubricator, Fig. 407, 
under a pressure of about 25 pounds in excess of the steam pressure. 
Referring to Fig. 414, diaphragm valve D, in the bottom of the lubri- 
cator, is kept closed by the steam pressure admitted through pipes B. 
Thus the inlet pressure must be greater than that of the steam before 

the valve will open and admit oil 
to the engine. The oil, after enter- 
ing, passes upward through the sight- 
feed glass and downward through 
the hollow arm A to the steam pipe. 
The engine oil is forced by the 
pump to the various points under a 
pressure of about 20 pounds. The 
waste oil is caught in suitable re- 
ceptacles and, after being filtered, is 
returned to the storage tank by a 
steam pump. This pump is con- 
nected so that it can supply the 
storage tank either from the filter or 
with fresh oil from a large oil tank 
in the basement. By this arrangement all handling of oil in the engine 
room is done away with. 

387. Oil Filters. — After oil has been applied to machinery its 
lubricating properties become impaired on account of (1) contami- 
nation with anti-lubricating material, such as dust, metallic particles 
from wear, gum, acid, and resin; and (2) exposure to heat and the 
atmosphere which drives off part of the more volatile constituents and 
decreases the fluidity of the oil. 

In many small plants no attempt is made to reclaim oil that has 
once been used, since the quantity is so small that the cost and 
trouble involved would more than offset the gain. Where large quan- 
tities of oil are used, considerable saving may be effected by using 
it over and over again. To render the oil fit for reuse it must be 
thoroughly purified. The anti-lubricating matter is removed by pre- 
cipitation and filtration. 

Fig. 415 shows a section through a " White Star" oil filter and purifier. 
The apparatus consists of a cylindrical sheet-iron vessel divided into two 
compartments by a vertical partition. These two compartments are 
connected near the top by valve B. The smaller chamber is provided 
with a funnel A and a steam coil for heating the contents. The large 
chamber contains a cylindrical wire screen covered with several folds of 
filtering cloth. Impure oil is poured into funnel A, the upper part of 



688 



STEAM POWER PLANT ENGINEERING 




WATER LEVEL 



WATER '_-_ ~ 



which is provided with a removable sieve or strainer, and is discharged 
below the surface of the water through holes in the foot of the tube. 
The thin streams of oil rise vertically to the surface of the water and 
the heavy particles of grit and dirt gravitate to the bottom. The steam 

coil heats the oil and water 
and facilitates precipita- 
tion of the solid matter 
by thinning out the streams 
of oil. When the oil in the 
smaller chamber reaches the 
level of valve B it flows in- 
to the filter bag, which re- 
moves the remaining im- 
purities and permits the 
purified products to flow 
into the -large compartment 
from which it may be drawn 
at will. All parts are access- 
ible and readily removed 
for cleaning purposes. The 
accumulated sediment in the bottom of the small chamber is dis- 
charged to waste at intervals by means of a suitable drain. When the 



PERFORATED PLATE 
FILTERING MATERIAL 
PERFORATED PLATE 

PERFORATED PLATE 

FILTERING MATERIAL 

PERFORATED PLATE 
WATER 

STEAM COILS 



Fig. 415. White Star Oil Filter. 




SECTION 1 SECTION 2 SECTION 3 

Fig. 416. Turner Oil Filter. 



SECTION 4 



filter cloth is to be removed, valve B is closed and the wire cylinder is 
disconnected and lifted out. Any oil remaining in the filter is returned 
to funnel A. The filter cloth is held against the screen by cords and 
hence is readily removed. 



LUBRICANTS AND LUBRICATION 689 

Fig. 416 shows a section through a Turner oil filter, illustrating the 
type of filter usually installed in large stations where continuous fil- 
tration is desired. This apparatus consists of a rectangular tank 
divided into four compartments. The returns from the lubricating 
system flow into section 1 through a screened funnel and discharge 
into the water space at the bottom of the compartment. The oil rises 
through the water, passes, under pressure of the head in the funnel, 
through a layer of filtering material resting on a perforated plate, and 
collects in an inverted cone. Through perforations round the top of 
the cone it passes into a dirt chamber, where most of the heavy impuri- 
ties are deposited, and then, still rising, passes through another per- 
forated plate and more filtering material. The partially cleaned oil, 
which issues, overflows into the second compartment and thence into 
the third, the same cycle of operations being repeated in these two. 
The overflow from the third compartment descends through a final 
filter in the fourth compartment and collects at the bottom, from 
which it is withdrawn by the oil pump. 

Forced-Feed Lubrication: Am. Elecn., Aug., 1902, p. 402, Dec, 1905, p. 608; 
Automobile, Nov. 1, 1906, p. 572; Mech. Engr., April 20, 1907, p. 552. 

Cylinder Lubrication: Power, Dec, 1902, p. 30, Jan., 1905, p. 36, March, 1906, 
p. 163; Engr., Lond., 1905, Vol. 96, pp. 55, 108, 132, 155; St. Ry. Jour., June 22, 1907, 
p. 1103; Engr. U.S., Oct. 15, 1906, p. 682; Am. Gas Light Jour., Jan. 23, 1905, p. 130; 
Horseless Age, Sept. 24, 1902, p. 676. 

Miscellaneous. — Measurement of Durability of Lubricants: Trans. A.S.M.E., 
11-1013. Valuation of Lubricant by Consumer: Trans. A.S.M.E., 6-437. Suit- 
ability of Lubricants: Power, Nov., 1906, p. 673. Oil Required for Lubricators: 
Elec World, May 5, 1906, p. 934. Gumming Tests: Jour. Am. Chem. Soc, April, 
1902, p. 467. Valuation of Lubricants: Jour. Am. Chem. Ind., April 15, 1905, 
p. 315. 

Lubrication, General: Power, March, 1903, p. 135; Mech. Engr., June 30, 1906, 
p. 919; Prac Engr., Dec. 15, 1905, p. 915; Elec Engr., Lond., Sept. 7, 1906, p. 344. 

■ Oil Purification: Elec. Engr., Lond., Jan. 13, 1903, p. 51; Elec. World, Dec. 
1, 1906, p. 1053. 

Economy in Lubrication of Machinery: Trans. A.S.M.E., 4-315. Theory of 
Finance of Lubrication: Trans. A.S.M.E., 6-437. 

Experiments, Formulas, and Constants for Lubrication of Bearings; Am. Mach., 
Sept. 10, 1903, pp. 1281, 1316, 1350. 

Lubricators and Lubricants: Power & Engr., Sept. 21, 1909, p. 486. j 

Selection of an Oil for Lubrication: Power & Engr., July 27, 1909, p. 137. 



CHAPTER XVII. 

FINANCE AND ECONOMICS— COST OF POWER. 

388. Records. — Few engineers realize the importance of a detailed 
system of accounting, or the saving which may be effected in cost of 
operation by careful study of the daily records of performance. Many 
regard graphical load curves, meter readings, and similar records as 
interesting but of little economic value. During the past few years 
the author has made a close study of the cost of power in a large num- 
ber of central and isolated stations in Chicago, and found, without 
exception, that the highest economy was effected by the engineers 
who kept the most systematic records; the poorest results were 
obtained where records were kept indifferently or not at all. In some 
small plants the numerous duties of the engineer prevented him from 
devoting the necessary time, but in the majority of cases the absence 
of records was due entirely to lack of interest. Power-plant records to 
be of value must be closely studied with a view to improvement. The 
mere accumulation of data to be filed away and never again referred 
to is a waste of time and money. 

Records should cover not only the daily operation of the plant but 
also, as permanent statistics, a complete analysis of each item of 
equipment. The value of such data cannot be overestimated. The 
engineer will frequently find it greatly to his interest to have avail- 
able at a moment's notice the complete details of his engines, boilers, 
generators, and other machinery, especially when it is required to 
renew a broken or worn-out part. 

389. Output. — The periodical output of a power plant may be 
expressed in terms of 

(1) Steam plant. 

Indicated or brake horse power. 
Indicated or brake horse-power hours. 

(2) Electric lighting plant. 

Electrical horse power or kilowatts. 

Electrical horse-power hours or kilowatt hours. 

Lamp hours. 

690 



FINANCE AND ECONOMICS — COST OF POWER. 691 

(3) Electric railway plant. 
Electrical horse power or kilowatts. 
Electrical horse-power hours or kilowatt hours. 
Car miles. 

When a plant is operating at practically constant load it is suffi- 
ciently accurate for most purposes to express the output in horse 
power or kilowatts per year. When the output fluctuates from day 
to day it is best expressed in horse-power hours or kilowatt hours, or 
by specifying the load factor along with the periodical output in 
horse power. For example, 1 horse power per year, 24 hours per day 
and 365 days per year, is equivalent to 365 X 24 = 8760 horse-power 
hours. If the full power is used throughout this time, it matters little 
whether the charge is based on horse power or horse-power hours; if, 
however, the power is used say only half the time, the yearly cost per 
horse power will remain unchanged but the cost per horse-power hour 
will be just double. As will be shown later the load factor (ratio of 
actual to rated load) exerts a marked influence on the cost of pro- 
ducing power, and for this reason the output is usually expressed as 
horse-power hours, kilowatt hours, lamp hours, or the like. 

390. Load Factor. — The yearly load factor or simply load factor, as 
it is usually called, is the ratio of the actual yearly output to the rated 
yearly output measured on a 24-hour basis. Thus: 

For a steam plant, 

Load factor = Yearly output, horse-power hours . ( 

Rated horse power X 8760 

For an electric station, 

Load factor = Yearly output kilowatt hours _ ( 

Rated capacity, kilowatts, X 8760 
(8760 = number of hours in one year.) 

The curve load factor or station load factor is the ratio of the actual 
yearly output to the rated output, based upon the number of hours the 
plant is in actual operation. Thus for an electric station, 

Curve load factor- Yearly output, kilowatt hours _ ( 

Rated capacity X hours plant is in operation 

In any plant the great desideratum is a high load factor with con- 
sequent greatest return on the investment. All the factors of expense 
included in the cost of power are then operating at maximum economy. 
High peak loads and low average loads necessitate large machines 
which are but little used and greatly increase the fixed charges. 



692 



STEAM POWER PLANT ENGINEERING 



In any system the total fixed charges per year are constant irre- 
spective of the load factor, since interest, taxes, depreciation, insurance, 
and maintenance go on whether the plant is in operation or not. The 
total fixed charges for a specific case are illustrated in Fig. 417 by a 















































1.4 










































280000 
































cfr*^ 








1.2 
u 

3 












S V6 


s* 














. v*> 


P°> 










240000 














\^c 


* 










^ 


3^ 














O 

w 

& 1.0 
















\V 


& 




^2^ 


















L, Dollar 








































O 

^ 0.8 
























<^ 


L* 


v ^i 














1 8 
tal Yearly Cos 






















J# 


















2 

g 0.6 

o 


















^< 


Ay 


Oper v tij 




ost; 




























«*?> 








Cer 


ts PerKw. 


Hr. 






kp^ 
















8 

80000 


o 

O 04 










<5 


0^ 




































X 


S 






< — Total Fixed Charges, Dolls 


irs 












s 












>• 














2 
























--i£ 


^C 


harg 


. e « J 


>er 


Kfl. 


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40000 








































uts 




0.0 

< 













































) 


1 





2 





3 



Ye 


4 
arly 




L0£ 


5 
idF 



acto 


G 
r-P 



erC 


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ent 





£ 





9 





1C 






Fig. 417. Influence of Load Factor on the Cost of Power at the Switchboard. (5000 
Kilowatt Electric Light and Power Station.) 

straight line. The cost per kilowatt hour, however, decreases as the 
load factor increases. For example, with the plant operating con- 
tinuously at rated load (100 per cent load factor) the fixed charges 
per kilowatt hour are 

— ^955 — = $0.00148. 
5000 X 8760 

With 30 per cent load factor these charges are 

65 ' 0QQ = $0.00445 kilowatt hour. 

0.3 (5000 X 8760) 

The higher the load-factor the greater is the amount of power produced 
and the longer does the apparatus work at best efficiency. But the 
greater the power produced the larger will be the fuel consumption 
and the oil and supply requirements. The labor charges will be prac- 
tically constant. The total operating cost per year increases as the 
load factor increases, but not directly. (See Fig. 417.) The cost per 



FINANCE AND ECONOMICS — COST OF POWER 693 

kilowatt hour, however, decreases as the load factor increases. For 
example, the operating costs per year with plant operating contin- 
uously at full load are $230,200. This gives 

230 ' 2QQ — = $0.00525 per kilowatt hour. 
5000 X 8760 F 

With 30 per cent load factor the yearly operating charges are 
$87,980, which gives 

, 87,98Q = $0.0067 per kilowatt hour. 

0.3 (5000 X 8760) F 

Table 107 shows the influence of the load factor on the cost of 
power in two isolated stations of the same rated capacity, one operat- 
ing with the unusually high load factor of 80 per cent and the other 
operating with the low load factor of 17 per cent. The former fur- 
nishes current for a large electro-chemical concern in which the load 
is practically constant. 

In general, the higher the load factor the greater becomes the ratio 
of the operating to the fixed charges, and extra investment may become 
advisable to secure the greatest economy possible. 

On the other hand, when the load factor is low the fixed charges are 
the governing factor in the cost of power, and extra expenditures 
must be carefully considered, particularly if fuel is cheap. 

391. Cost of Operation. — The cost of operation of power plants is 
conveniently divided into two parts: 

(1) Fixed charges. 

(a) Investment costs. 

(b) Administration costs. 

(2) Operating costs. 

392. Fixed Charges. — These cover all expenses which do not 
expand and contract with the output. In very large plants they are 
usually divided into two parts, (a) the investment costs, which include 
interest, rental, depreciation, taxes, and insurance, and a reserve fund 
to cover depreciation of the investment, and (b) the administration 
costs, which include rental of offices, annual salaries of officers, and all 
other expenses not directly chargeable to the power plant. In the 
average plant the fixed charges comprise interest, rental, depreciation, 
taxes, insurance, and sometimes maintenance, though the latter is 
ordinarily included in the operating costs. 

393. Interest. — The rates of interest on borrowed money vary 
with the nature of the security. In the case of power plants the form 
of security is usually a mortgage on the plant and equipment. If a 



694 STEAM POWER PLANT ENGINEERING 

builder has sufficient funds to construct the plant without borrowing, 
he should charge against the item " interest " the income which the sum 
involved would bring if placed out at interest or if invested in his own 
business. In estimating the interest charges 5 per cent of the capital 
invested is ordinarily assumed unless specific figures are available. 

TABLE 90. 

APPROXIMATE USEFUL LIFE OF VARIOUS PORTIONS OF STEAM POWER PLANT 

EQUIPMENTS. 

Years. 

Buildings, brick or concrete 50 

Buildings, wooden or sheet iron 15 

Chimneys, brick '. : 50 

Chimneys, self-sustaining steel 25 

Chimneys, guyed sheet-iron 10 

Boilers, water-tube 25 

Boilers, fire-tube 15 

Engines, slow-speed 25 

Engines, high-speed 15 

Turbines 25 

Generators, direct-current 25 

Generators, alternating-current 30 

Motors 20 

Pumps 25 

Condensers, jet 35 

Condensers, surface 20 

Heaters, open 30 

Heaters, closed 20 

Economizers 20 

Wiring 20 

Belts 7 

Coal conveyor, bucket 15 

Coal conveyor, belt 10 

Transformers, stationary 30 

Rotary converters 25 

Storage batteries 15 

Piping, ordinary 12 

Piping, first class 20 

NOTE. — So much depends upon the design and the conditions of operation that no fixed 
values can be definitely assigned and the above figures should be used with caution. Practice 
shows that most power-plant appliances become obsolete long before the limit of their useful 
life is reached. 

394. Depreciation. — This charge represents the gradual deterio- 
ration of a plant, resulting in its eventually wearing out. It is also 
assumed to represent the superannuation of a plant or the rate at 
which the apparatus is becoming obsolete. Thus, under the first 
assumption, if the useful life of an engine is 40 years, the rate of 
depreciation, neglecting interest, is 2.5 per cent; if, however, it is 
assumed that the engine will become obsolete in 20 years and uneco- 
nomical for further operation, the rate of depreciation will be 5 per 
cent. It is difficult to assign a fixed rate of depreciation against any 



FINANCE AND ECONOMICS — COST OF POWER 



695 



piece of apparatus, due to possible new developments which cannot 
be reckoned with in advance in computing the useful life of the appa- 
ratus. Again, depreciation cannot always be separated from current 
repairs and is a variable factor even in the parts of the same machine. 
It is therefore more or less of an approximation. The average life of 
various parts of a steam power plant is outlined in Table 90, but 
on account of the inability to assign fixed values to the useful life of 
any apparatus, and on account of the great number of appliances in 
even a small plant, it is customary to charge a fixed rate of depreciation 
against the entire plant and thus avoid confusion and complexity. 
This very crude method usually results in overestimation in well- 
designed, well- operated plants and underestimation in poorly designed 
and badly managed installations. One of the largest power plant 
designing concerns in Chicago charges 7J per cent against deprecia- 
tion and finds this figure none too small. The Pennsylvania Railroad 
uses 7 per cent to cover depreciation charges on their power-house 
equipments. 

TABLE 91. 

RATE OF DEPRECIATION. 



(Per Cent of First Cost.) 







Rate of Interest, per Cent. 




2 


2.5 


3 


8.5 


4 


4.5 


5 


5.5 


6 


7 


8 


9 


10 




2 


49.50 


49.37 


49.27 


49.14 


49.02 


48.90 


48.78 


48.66 


48.54 


48.31 


48.07 


47.84 


47.62 




3 


32.67 


32.51 


32.35 


32.19 


32.03 


31.87 


31.72 


31.56 


31.41 


31.10 


30.80 


30.51 


30.21 




4 


24.26 


24.08 


23.90 


23.72 


23.55 


23.39 


23.20 


23.03 


22.86 


22.52 


22.19 


21.84 


21.55 




5 


19.21 


19.02 


18.83 


18.65 


18.46 


18.28 


18.10 


17.91 


17.73 


17.40 


17.04 


16.73 


16.37 




6 


15.85 


15.65 


15.46 


15.26 


15.08 


14.89 


14.70 


14.52 


14.33 


13.97 


13.63 


13.29 


12.96 


B9 


7 


13.45 


13.25 


13.05 


12.85 


12.66 


12.46 


12.28 


12.09 


11.91 


11.15 


11.20 


10.87 


10.55 


3 


8 


11.65 


11.44 


11.24 


11.05 


10.85 


10.66 


10.47 


10.28 


10.10 


9.74 


9.40 


9.06 


8.74 


3 


9 


10.25 


10.04 


9.84 


9.64 


9.45 


9.26 


9.07 


8.88 


8.70 


8.34 


8.00 


7.68 


7.36 




10 


9.13 


8.92 


8.72 


8.52 


8.33 


8.14 


7.95 


7.76 


7.58 


7.23 


6.90 


6.58 


6.27 


< 


11 


8.21 


8.01 


7.80 


7.61 


7.41 


7.22 


7.04 


6.85 


6.68 


6.33 


6.00 


5.69 


5.40 


o 


12 


7.45 


7.25 


7.04 


6.85 


6.65 


6.46 


6.28 


6.10 


5.92 


5.60 


5.27 


4.97 


4.69 


<u 


13 


6.81 


6.60 


6.40 


6.20 


6.01 


5.83 


5.64 


5.47 


5.29 


4.96 


4.65 


4.36 


4.08 


3 


14 


6.26 


6.05 


5.85 


5.65 


5.46 


5.28 


5.10 


4.93 


4.75 


4.49 


4.13 


3.84 


3.58 


3 


15 


5.78 


5.57 


5.37 


5.18 


4.99 


4.81 


4.63 


4.46 


4.29 


3.97 


3.68 


3.40 


3.15 


V 


16 


5.36 


5.16 


4.96 


4.77 


4.58 


4.40 


4.22 


4.06 


3.89 


3.58 


3.30 


3.03 


2.78 


-3 


17 


4.99 


4.79 


4.59 


4.40 


4.22 


4.04 


3.87 


3.70 


3.54 


3.24 


2.96 


2.71 


2.47 


3 


18 


4.67 


4.46 


4.27 


4.08 


3.90 


3.72 


3.55 


3.39 


3.23 


2.94 


2.66 


2.42 


2.19 


a 


19 


4.37 


4.17 


3.98 


3.79 


3.61 


3.44 


3.27 


3.11 


2.96 


2.67 


2.47 


2.17 


1.95 


DO 


20 


4.11 


3.91 


3.72 


3.53 


3.36 


3.19 


3.02 


2.87 


2.71 


2.44 


2.18 


1.95 


1.95 




25 


3.12 


2.92 


2.74 


2.56 


2.40 


2.24 


2.09 


1.95 


1.82 


1.58 


1.36 


1.18 


1.75 




30 


2.46 


2.27 


2.10 


1.93 


1.78 


1.64 


1.50 


1.38 


1.26 


1.06 


0.88 


0.73 


0.61 




35 


2.00 


1.82 


1.65 


1.50 


1.36 


1.23 


1.10 


0.99 


0.89 


0.72 


0.58 


0.46 


0.37 




40 


1.65 


1.48 


1.32 


1.18 


1.05 


0.93 


0.83 


0.73 


0.64 


0.50 


0.38 


0.29 


0.22 




45 


1.39 


1.22 


1.07 


0.94 


0.82 


0.72 


0.62 


0.54 


0.47 


0.35 


0.26 


0.19 


0.14 




50 


1.18 


1.02 


0.88 


0.76 


0.65 


0.56 


0.42 


0.40 


0.34 


0.25 


0.17 


0.12 


0.09 



696 STEAM POWER PLANT ENGINEERING 

The rate of depreciation in terms of interest and useful life is a 
simple problem in compound interest, and may be expressed: 

in which (1 +r) n -l 

d = rate of depreciation, per cent of first cost. 

r = rate of interest. 

n = assumed life of the apparatus in years. 

This is based on the assumption that the interest is compounded 
annually and that the apparatus is valueless at the end of n years. 
Table 91 has been calculated with this formula and gives the rate of 
depreciation for different rates of interest and different asssumptions 
as to useful life of apparatus. 

TABLE 92. 

DEPRECIATION PERCENTAGES DETERMINED BY THE TRACTION VALUATION 

COMMISSION. POWER-PLANT DEPRECIATION. 

Chicago, 111., Sept. 8, 1906. Per Cent. 

Engines, Corliss, slow-speed 3 to 5 

Engines, automatic, high-speed 5 to 10 

Cable-winding machinery 3 

Generators, direct connected, modern 5 

Generators, belted (depending on date) 5 to 10 

Traveling cranes 2 

Switchboard and all wiring 2 

Piping 35 

Pumps 5 

Heaters, closed 6 to 10 

Heaters, open, if cast iron only 3 

Breeching and connections, brick 5 

Breeching and connections, steel 10 

Boilers and settings, horizontal tubular 10 

Boilers and settings, water- tube 3.5 

Grates 10 

Stokers See below 

Coal-handling machinery 6 

Ash-handling machinery 8 

Combined coal and ash-handling machinery 7 

Storage bins, steel 3 to 10 

Miscellaneous items 5 

The above annual rates of depreciation have been used as a basis in depreciating 
the power-plant equipments. Apparatus has been depreciated at these rates down 
to 20 per cent of the wearing value, the wearing value being determined by sub- 
tracting the scrap value from the cost new. All power-plant equipment has been 
considered as worth 20 per cent of its wearing value as long as it is in operating 
condition. Depreciation applied to wearing value, as apparatus is always worth 
scrap value. 

Stokers. The fixed parts depreciate very little and the moving parts and grates 
very rapidly, as the moving parts are renewed and maintained in good condition. 
All stokers in operation have been depreciated 25 per cent. 

The above percentages applied to a particular plant of 3900 kilowatts capacity 
give an approximate depreciation for the whole plant of 4 per cent. 



FINANCE AND ECONOMICS — COST OF POWER 697 

Table 92 gives the depreciation percentages determined by the 
Traction Valuation Commission, Chicago, 111., as reported by the com- 
mission Sept. 18, 1906. 

Example: A 1000-square-feet surface condenser and auxiliaries cost 
$3500. With interest at 5 per cent, required the rate of depreciation, 
assuming a life of 25 years. 

d = 10 ° (1 + O.OsW x = 2.09 per cent. 

That is to say, if 2.09 per cent of the first cost is laid aside each year 
for 25 years at 5 per cent interest, compounded annually, it will equal 
the first cost at the end of this period. The sum thus laid aside 
is sometimes called the sinking fund. The solution is more readily 
obtained with the aid of Table 91; e.g., at the intersection of vertical 
column 5 (interest) and horizontal column 25 (life in years) we find 
the depreciation 2.09 per cent. This sinking-fund method of rating 
the depreciation is peculiarly adapted to power-plant practice, inas- 
much as a sum set aside at comparatively low rates of interest and 
compounded increases very slowly at first but more and more rapidly 
from year to year. This is precisely what happens in the deterioration 
of the plant. The loss of value is slight at first when the materials 
are new and usefulness is at a maximum, while towards the end of 
life both value and usefulness decline very rapidly. 

If the apparatus still has some value at the end of n years and if b is 
the ratio of the value of the old material to that of the new, the rate 
of depreciation becomes 

d = ioo ,/ (1 7 &) • (174) 

(1 + r) n — 1 

Example : In the foregoing problem, required the rate of depreciation 
if the value of the condenser outfit is $350 at the end of 25 years. 

Here b = — = .1. 
3500 

d = 100 X 0.05 (1-0.1) 
(1 + .05) 25 -l 

= 1.97 per cent. 

That is, 1.97 per cent of $3500, or $68.95, laid aside each year for 
25 years at 5 per cent interest and compounded annually will equal 
$3500 - 350, or $3150, at the end of this period. 



698 STEAM POWER PLANT ENGINEERING 

It is not supposed that an owner will regularly lay aside this sum 
annually, or take the trouble to arrange for its investment at current 
rates in the market or savings bank, since the money is probably worth 
more to him in his own business. In practice it is retained in his 
business or investments and is earning the rate of interest obtainable 
therein, but in determining the net profit or loss this depreciation item 
is nevertheless accounted for just as if it were actually placed in out- 
side investments. 

In appraising the present value of any apparatus in terms of the rate 
of interest and useful life the expression becomes 

V = 100 (J +r) m -l ( 

(1 +r )n_l ' 

in which 

V = total depreciation, per cent of original cost. 

m = number of years apparatus has been in operation. 

n = assumed life of apparatus. 

r = rate of interest. 

Example: In the preceding example, required the present valuation 
of the condenser, assuming that it has been in use 15 years. 

m = 15, n = 25, r = 0.05. 

Substituting these values in (175), 

ir inn (1 + 0.05) 15 - 1 AK - 

V= 100 i — ! '— = 45.1 per cent. 

(1 + 0.05) 25 ~1 F 

That is, the condenser has depreciated 45.1 per cent of its original 
value and consequently is worth $3500 — 0.451 X 3500 = $1921.50. 

Table 91 may be conveniently used in this connection. At the 
intersection of vertical column 5 and horizontal columns 15 and 25 
we find 4.63 and 2.09 respectively. Dividing 2.09 by 4.63 we get 
0.451 = 45.1 per cent, the total depreciation. Depreciation is often 
taken care of under the different items pertaining to maintenance, and 
whenever a change or repair is necessary it is charged directly into 
expense as maintenance. 

Though usually considered separately, interest and depreciation are 
sometimes considered as a single item, and in this case the rate of 
depreciation represents the sum which must be laid aside each 
year for the eventual renewal of apparatus plus the interest on the 
investment. 



FINANCE AND ECONOMICS - COST OF POWER 699 

395. Maintenance. — Maintenance usually refers to the expense of 
keeping the plant in running order over and above the cost of attend- 
ance. It includes cost of upkeep, replacement, and precautionary 
measures. This latter item includes the renewal of working parts, 
painting of perishable or exposed material, and replacing worn-out and 
defective material. Many engineers make no allowance for mainte- 
nance in the fixed charges and include these costs under supplies, 
attendance, or repairs. In a general way, when maintenance is 
included under the fixed charges, an annual charge of 2 per cent is con- 
sidered a liberal allowance, since most of the repair work comes under 
attendance. In street-railway practice maintenance is divided among 
the several parts of the system as follows: Buildings, steam appli- 
ances, electrical equipment, and miscellaneous. In this connection the 
maintenance becomes a part of the fixed charges, since the various 
items vary widely from month to month. 

396. Taxes and Insurance. — Taxes vary from a fraction of one per 
cent to one and one-half per cent, depending upon the location of the 
plant. An average figure is one per cent of the actual value of the 
investment. Buildings and machinery are ordinarily insured against 
fire loss and boilers against accidental explosions, and accident policies 
are sometimes carried on all operating machinery. A fair charge for 
this item is one-half per cent. 

397. Operating Costs. — Operating costs are conveniently divided as 
follows : 

(1) Labor or attendance. 

(2) Fuel and water. 

(3) Oil, waste, and supplies. 

(4) Repairs and maintenance. 

In large stations it is often desirable to keep the expenses of the 
various departments separate from those of the power plant proper. 
Thus in central stations the distributing system is an important 
branch and the attending expenses form a considerable portion of the 
total. They are therefore kept separate, since they are not strictly 
chargeable to power generation. In isolated stations the wages of 
elevator men, though in a general way a part of the power-plant 
expenses, are not included in the " labor and attendance " charge of 
the plant. Lamps are a large item of expense in a tall office build- 
ing, and for this reason are often kept separate from supplies. 

398. Labor, Attendance, Wages. — The minimum number of men 
required to handle a given plant is approximately a fixed quantity 
and it is seldom possible to so arrange the work that any material 



700 



STEAM POWER PLANT ENGINEERING 



reduction can be effected. Until very recently it has been the uni- 
versal custom to pay wages ona" flat rate " basis, that is, the attend- 
ant is given a fixed sum per day or month irrespective of the amount 
of work required or the economy of operation. In many cases, how- 
ever, the bonus system has been successfully adopted. For example, 
in the boiler room the coal consumption is determined for a given 
period of time with ordinary careful firing, and the fireman is offered 
a reasonable percentage on the saving of coal which he is able to 
effect over this record by special care and attention to the keeping of 
fires always in the best condition, avoiding the blowing off of steam, 
using as little coal as needed for banking fires, and in other ways. 
Where careful records are kept of supplies, repairs, and renewals, the 
bonus is also applicable to electricians, oilers, and other employees. 

Labor should always be estimated or recorded as so many dollars 
per month or per year and not merely in terms of the output unless 
the load factor is definitely known, otherwise comparisons are mis- 
leading. For example, consider two plants of 500 kilowatts capacity, 
each with labor charges, say, of $400 per month. Suppose the output 
of one is 100,000 kilowatt hours per month and that of the other 
40,000 kilowatt hours per month. The monthly charges are evidently 
the same, viz., $400, but the cost per kilowatt hour differs widely, 
being 0.4 cent in the first case and 1 cent in the latter. 

The cost of labor varies so much with the location of the plant and 
the conditions of operation that general figures are of little value 
except as a rough guide. The figures in Table 93 and Table 94 give 
average results for general practice. Specific figures will be found in 
Tables 95 to 107. 

399. Cost of Fuel. — Tables 95 to 107 give examples of the cost of 
fuel in different types and sizes of steam power plants. It will be 
noted that this item varies considerably even with plants of the same 
general class. So much depends upon the market price of the fuel 
itself that the item " cost per horse-power or kilowatt hour " gives 
little information concerning the economy of operation unless the 
price of the fuel, its calorific value, and the water rate of the prime 
movers are specified. In a general sense the cost of fuel will range 
from 40 to 70 per cent of the total operating expenses. In estimating 
the cost of fuel for a proposed installation due consideration should be 
given to the coal burned in banking fires, heat lost in blowing off 
boilers, and reduced efficiency in operating at underloads and over- 
loads. For example, individual tests of a number of boilers in a large 
central station in Chicago gave an average evaporation of 6.1 pounds 
of water per pound of coal, actual conditions, whereas the evaporation 



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FINANCE AND ECONOMICS — COST OF POWER 703 

determined by dividing the total water fed into the boiler per year 
by the total consumption of coal gave only 5.2 pounds. Current 
practice gives an average efficiency (based on yearly operation) of 
boiler and furnace of 70 per cent for pumping stations running at 
practically full load, 65 per cent for large lighting and power stations 
with yearly load factor of 0.50 or more, and 60 per cent for similar 
stations with load factor between 0.35 and 0.45. For very low load 
factors, 0.25 and under (as in connection with manufacturing plants, 
tall office building and other plants operating on a 10-hour basis) 
this efficiency seldom exceeds 50 per cent. With these figures as a 
guide the cost of fuel per unit output may be roughly approximated. 

400. Oil, Waste, and Supplies. — These items approximate 2 to 10 
per cent of the total operating expenses. Tables 95 to 107 give some 
idea of current practice in different classes of power plants. 

401. Repairs and Maintenance. — This item ordinarily refers to the 
cost of keeping the plant in running order over and above the cost of 
labor or attendance, and depends upon the age and condition of the 
plant and the efficiency of the employees. Tables 95 to 107 give the 
cost of repairs and maintenance for a wide range in power-plant 
practice. 

402. Cost of Power. — The actual cost of producing power depends 
upon the geographical location of the plant, the size of apparatus, the 
design, conditions of loading, system of distribution, and the method 
of accounting. Tables 95 to 108 compiled from various sources give 
the detailed costs of a large number of central and isolated stations. 

Table 95. Operating costs per kilowatt hour for a number of typical 
British electric light and power plants. 

Table 96. Operating costs per kilowatt hour for a number of United 
States electric power plants for street-railway, light, and general power 
service. 

Table 97. Average operating costs per kilowatt hour for all stations 
of the Boston Elevated. 

Table 98. Operating costs for the year 1907 of the mechanical plant 
of the First National Bank Building, Chicago. 

Table 99. Costs, fixed and operating, of producing one brake horse 
power per year, simple non-condensing engine, etc. 

Table 100. Cost of power for compound condensing engine plants. 

Table 101. Costs of different sizes and types of plants and annual 
costs per brake horse power, average working conditions. 

Tables 102, 103, and 104. Cost, fixed and operating, of producing 
one electrical horse power per year for different sizes and types of 
plants. 



704 



STEAM POWER PLANT ENGINEERING 



Table 105. Influence of load factor on cost of electrical power in 
isolated stations. 

Tables 106 and 107. Cost of furnishing heat, light, and power for a 
number of isolated stations in New York City, tall office buildings, loft 
buildings, apartment houses, hotels, and club buildings. 



135 
130 
125 
120 
115 

03 

| 110 
pq 

£ 105 

o 

W 100 



90 







































^ 










































































































Curves showing Range 
in Cost of Power 
in 200 Mfg. Plants 

Middle-Western States 




1 


































































\ 










































[ 








































\ 








































\ 








































\ 










































\ 








































\ 




































\ 


I 




\ 




































\ 




\ 




































\ 








































\ 
































































































^ 
































/, 0l 










;& 






























•flyj 


' 
















































































































llllllil 

Size of Plant, Horse -Power 
Fig. 418. 



o o 

8 g 



Table 108. Cost of furnishing heat, light, and power for the year 
1907, First National Bank Building, Chicago. 



FINANCE AND ECONOMICS — COST OF POWER 




705 



20 40 60 

PER CENT LOAD-FACTOR 



Fig. 418a. Cost of Power in Large Power Plants with Maximum Load over 
30,000 Kilowatts. Coal at $3.00 per Ton. 14,500 B.T.U. per Pound. 



706 



STEAM POWER PLANT ENGINEERING 



130 



50 

























l.=Keciprocating steam-plant. 

2.=- Steam-turbine plant. 

3. -Reciprocating-engine and low-pressure 

turbine plant. 
4. = Gas-engine plant. 
5 = Gas-engine and steam-turbine plant. 
6 . = Hydraulic-plant 










\y 




























































2^ 




















5^ 




























^ 


^ 


^ 










1F^ 

^2 















































































20 



Per Cent Load-Factor 



Fig. 418b. Cost of Power in Large Power Plants with Maximum Load over 30,000 
kilowatts. Coal at $1.50 per Ton. 11,000 B.T.U. per Pound. 



FINANCE AND ECONOMICS — COST OF POWER 



707 



Table 94a, and Figs. 418a, 418b and 418c give the fundamental 
relations between the various items entering into the cost of power for 
various types of plants of over 30,000 kilowatts rated capacity. These 
data are taken from a paper presented by H. G. Stott at a meeting of 
the Toronto section of the American Institute of Electrical Engineers, 
Toronto, Ont., December, 1908. The figures have been brought up to 
date (June, 1910) by Mr. Stott and show what is actually being done 
to-day in large plants of the size stated above. 

TABLE 94a. 

DISTRIBUTION OF MAINTENANCE AND OPERATION COSTS IN POWER PLANTS 
HAVING A MAXIMUM LOAD OVER 30,000 KILOWATTS. . 

(H. G. Stott.) 





Recip- 
rocating 
Steam 
Plant. 


Steam 
Tur- 
bine 

Plant. 


Recip- 
rocating 
Engines 
and Low- 
pressure 
Steam 
Tur- 
bines. 


Gas 

En- 
gine 
Plant. 


Gas 
En- 
gines 
and 
Steam 
Tur- 
bines. 


Hy- 
drau- 
lic. 


Maintenance. 

1. Engine room, mechanical 

2. Boiler or producer room 


2.59 
4.65 
0.58 
1.13 

61.70 
7.20 
6.75 
7.20 
2.28 
1.07 
2.54 
1.78 
0.30 
0.17 

100.00 

100.00 

125.00 

11% 


0.51 
4.33 
0.54 
1.13 

55.53 
0.65 
1.36 
6.74 
2.13 
0.95 
2.54 
0.35 
0.30 
0.17 

77.23 

75.00 

93.75 

11% 


1.55 
3.55 
0.44 
1.13 

46.48 
0.61 
4.06 
5.50 
1.75 
0.81 
2.54 
1.02 
0.30 
0.17 

69.91 

80.00 

100.00 

11% 


5.18 
1.16 
0.29 
1.13 

26.52 
3.60 
6.76 
1.81 
1.14 
0.54 
2.54 
1.80 
0.30 
0.17 

52.94 
110.00 
137.50 

12% 


2.84 
1.97 
0.29 
1.13 

25.97 

2.16 
4.06 
3.05 
1.14 
0.54 
2.54 
1.07 
0.30 
0.17 

47.23 

96.20 

120.00 

11.5% 


0.51 


3. Coal- and ash-handling apparatus. 

4. Electrical apparatus 


1 13 


Operation. 

5. Coal 

6. Water 




7. Engine room, labor 


1.36 


8. Boiler or producer room, labor . . . 

9. Coal- and ash-handling, labor .... 
10. Ash removal 




11. Electrical labor 


2.54 


12. Engine room, lubrication 

13. Engine room, waste, etc 

14. Boiler room, lubrication, etc 

Relative operating cost, per cent 

Relative investment, per cent ........ 

Probable average cost per kilowatt . . . 
Probable fixed charges 


0.20 
0.20 

5.94 
100.00 
125.00 
11% 





For steam-turbine plants larger than 60,000 kw. the cost per kilowatt may be 
reduced to $75.00. 



08 



STEAM POWER PLANT ENGINEERING 

















i 1 1 1 
FIXED CHARGES 

1 1 1 i 










10 






1. 


Reciprocating steam plant. Cost $125.00 per Kw. 

.1 I*' 1 J 






14 






2| 




" $75.00 4 » 














3. 


Eng 


no and low pressure 


turb 


nc plant. 










12 






4. 


Cost poo.uu per Kw. 

Gas engine plant. Cost $137.50 per Kw. 














5. 


Gas 


engine and steam turbine 


plant. 










10 


CO 
UJ 

o 


1\ 


6. 


Cost 
Hyd 


$15*0.00 per 
raulic plan 


Kw. 1 

t. Cost $125.00 


I 
per Kw. 










< 


llV 








1 1 
TOTAL OPERATING 


CHARGES 

| 






8 


o 

Q 


|\ 




1-6 A 

l'P' 


s above, coal (& $3.00- 

dittO «nn.lL<a SllfiO— 1 


-1450( 
1000- 


B.t.u. ix 
Bti™ 


'rib. 
lb 




UJ 


























— 


6 


































\\ \ \ 
























4 






WY 
























* 




X " ^0 




w 

g1\ 




















2 










»V^ 


^<3 


^t 




























2f*- 




^^^^^ 












































CO 
UJ 

C3 












































6" 


2 


CC 
< 






















_£$' 








O 
CD 




.,_ 





— Z." 


"ZTZ- 





ZZZ1 





e 


W~' 


^ 




L^2' 






s.-£2t"^Z\— — 












\A 






4 


1- 
— <- 


^S^ 


^-~ 




















f 3 


"V 






CC 
LU 


s 














-~' r 
















O 
_1 


tf?2 














K 1 


































6 


< 

Q 
































r- 































20 



40 



60 80 100 

PERCENT LOAD 



120 



140 



Fig. 418c. 



Cost of Power in Large Power Plant with Maximum Load over 
30,000 kilowatts. 



FINANCE AND ECONOMICS — COST OF POWER 



709 



TABLE 95. 

OPERATING COSTS, CENTS PER KILOWATT HOUR. TYPICAL BRITISH ELECTRIC 
LIGHT AND POWER PLANTS, 1902. 

(Engineering Record, March 26, 1904, p. 389.) 



Locality. 



Bradford. . . . 

Leeds , 

St. Helen's. . 
Edinburgh . . 

Bolton 

Booth 

Liverpool 
South Shields 
Nottingham. . 

Preston 

Farnsworth. . 
Leith 



Kilowatts 
Installed. 



6,380 

8,740 

1,340 

10,477 

3,700 

850 

21,190 

1,600 

5,642 

1,920 

610 

990 



Yearly- 
Load 
Factor. 



Per Cent. 
20.93 
12.31 
17.84 
14.75 
18.87 
28.44 
25.11 
15.82 
12.97 
13.31 
14.54 
19.79 



Coal. 



0.52 
0.56 
0.52 
0.68 
0.70 
0.82 
0.74 
0.74 
0.92 
0.72 
0.92 
1.10 



Oil, Waste, 




Re- 


and 


Wages. 


Supplies. 




pairs. 


0.10 


0.16 


0.26 


0.06 


0.34 


0.28 


0.06 


0.34 


0.38 


0.08 


0.18 


0.36 


0.12 


0.3 


0.20 


0.06 


0.3 


0.22 


0.12 


0.3 


0.26 


0.08 


0.4 


0.30 


0.20 


0.32 


0.18 


0.12 


0.36 


0.46 


0.20 


0.36 


0.22 


0.08 


0.42 


0.18 



Total. 



1.04 

1.24 

1.3 

1.3 

1.32 

1.4 

1.42 

1.52 

1.62 

1.66 

1.7 

1.78 



TABLE 96. 

OPERATING COSTS, CENTS PER KILOWATT HOUR. TYPICAL STREET-RAILWAY 
PLANTS, UNITED STATES. 

(Street Railway Review, Oct. 20, 1902, p. 774.) 





1 


2 


3 


4 


5 


Items. 


Two 

Stations 
Supplying 
Power for 
1000 Cars, 
175 Miles 

of Track. 
(Belted In- 
stallation.) 


One 

Station, 
Power and 
Licht, 500 

Cars, 
230 Miles 
of "Track. 


Large 

City 

Central 

Station. 


Combined Light 
and Power Station. 


Inter* 
urban, 

Light 
and 




July. 


November. 


Power. 


Fuel 


0.68 
0.199 
0.013 
0.04 

0.005 
0.072 

0.008 
0.033 


0.803 
0.133 
0.018 


0.509 

0.115 

0.012 

.009 


0.958 
0.287 
0.048 


0.870 
0.276 
0.025 


1.004 


Wages 


0.217 


Oil and waste 

Water 


0.023 
0.022 


Maintenance: 

Buildings 








0.015 


Steam appliances . 
Electric equip- 
ment 


0.029 

0.019 
0.023 


.011 

.001 
.012 






0.035 


0.004 




0.010 


Miscellaneous 














Total 


1.05 

1.55 

20,981,295 

8.8 


1.03 
2.66 
2,140,641 
6.04 


0.669 

3.07 

641,650 


1.299 

2.98 

119,304 

6.4 


1.188 

2.98 

114,384 

5.55 


1.326 


Cost of coal, dollars 
per ton 


3.62 


Output, kilowatt 
hours 


2,104,337 
5.55 


Coal per kilowatt 
hour, pounds 







710 



STEAM POWER PLANT ENGINEERING 



TABLE 97. 

OPERATING EXPENSES OF BOSTON ELEVATED RAILWAY COMPANY. 
AVERAGE OF ALL STATIONS AT SWITCHBOARD, APRIL 8, 1910. 





Year. 




1902 


1903 


1904 


1905 


1906 


1907 


1908 


1909 


Rated capacity, kilowatts 

Yearly load factor 

Coal.* Cents per K.W.H.. 
Labor. " " " 
Repairs.! " " " 
Supplies " " " . . 


35544 
.386 
0.47 
0.19 

0.58 


35544 
.391 
0.78 
0.18 

0.94 


35544 
.417 
0.53 
0.16 

0.67 


35544 
.437 
0.45 
0.15 

0.57 


38469 

.43 
0.47 
0.17 

0.60 


39969 

.45 
0.55 
0.18 

0.76 


50425 

.37 
0.56 
0.21 

0.86 


50063 

.37 

0.45 

0.20 

0.61 


Total " " . . 
Ratio operating expense 

to gross earnings 

Price of coal per ton 


0.77 

0.694 
$3.60 


1.12 

0.690 

$4,854 


0.83 

0.696 
$3.55 


0.72 

0.68 
$3.1354 


0.77 

0.687 
$3.1859 


0.94 

0.69 

$3,572 


1.07 

0.67 

$3,568 


0.81 

0.654 
$3,209 



Coal included in supplies. 



t Repairs included in supplies and labor. 



TABLE 98. 
COST OF OPERATION (1907). FIRST NATIONAL BANK BUILDING, CHICAGO. 



Coal bill 

•Ash cartage 

Water bill 

Electrical supplies and repairs 

Engine-room supplies and repairs 

Boiler-room supplies and repairs 

Oil, waste, and grease 

Packings 

Machine-shop supplies 

Refrigerating supplies 

Steam fitting 

Steam heat supplies 

Plumbing supplies 

Lamps 

Wages 

Petty expenses 

Office expenses 

Doctor bills ....:.... 

Coal analysis 

Total 

Total receipts from all sources for power, heat, light, and mis 
cellany 

Net cost 



$34,567.27 

2,075.83 

3,535.75 

4,016.68 

1,311.36 

1,316.50 

1,554.80 

727.01 

187.13 

761.81 

499.53 

110.20 

184.87 

1,655.22 

24,572.89 

24.62 

186.55 

43.00 

160.00 

$77,591.02 

71,435.85 

$6,055.17 



FINANCE AND ECONOMICS — COST OF POWER 



711 



TABLE 99. 

COST OF ONE HORSE POWER PER YEAR, SIMPLE ENGINES, NON-CONDENSING, 
10-HOUR BASIS, 308 DAYS PER YEAR. 

(Wm. O. Webber, Engineer U.S., March 16, 1903, p. 241.) 



Size of plant horse power 

Cost of plant per horse power 

Fixed charges at 14 per cent 

Coal per horse-power hour, in pounds. . . . 

Cost at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, and supplies 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



20 

$200.00 

28.00 

12.00 

66.00 

30.00 

6.00 

146.50 

138.25 

130.00 

121.75 

113.50 

105.25 

97.00 



40 

$190.00 

26.60 

10.00 

55.00 

20.00 

4.00 

119.35 

112.47 

105.60 

98.72 

91.85 

84.97 

78.10 



60 


180 


00 


25 


20 


9 


00 


49 


50 


15 


00 


3 


00 


105 


07 


98 


80 


92 


70 


86 


51 


80 


32 


74 


13 


67 


95 



80 
$175.00 
24.50 

8.00 
44.00 
13.00 

2.60 
95.10 
89.60 
84.10 
78.60 
73.10 
67.60 
■ 62.10 



TABLE 100. 

COST OF ONE HORSE POWER PER YEAR, COMPOUND CONDENSING ENGINES, 
10-HOUR BASIS, 308 DAYS PER YEAR. 

(Wm. O. Webber, Engineer U.S., Feb. 2, 1903, p. 144.) 



Size of plant horse power 

Cost of plant per horse power. . . . 

Fixed charges at 14 per cent 

Coal per horse-power hour, pounds 

Cost of fuel at $4.00 per ton 

Attendance, 10-hour basis 

Oil, waste, supplies 

Total 

With coal at $5.00 per ton 

With coal at $4.50 per ton 

With coal at $4.00 per ton 

With coal at $3.50 per ton 

With coal at $3.00 per ton 

With coal at $2.50 per ton 

With coal at $2.00 per ton 



100 


200 


300 


400 


500 


$170.00 


$146.00 


$126.00 


$110.00 


$96.00 


23.80 


24.40 


17.65 


15.40 


13.45 


7.0 


6.5 


6.0 


5.5 


5.0 


38.50 


35.70 


33.00 


32.00 


27.50 


12.00 


10.00 


8.60 


7.25 


6.20 


2.40 


2.00 


1.72 


1.45 


1.24 


76.70 


68.10 


60.97 


56.10 


48.39 


86.40 


77.10 


69.22 


61.90 


55.29 


81.50 


72.60 


65.07 


58.10 


51.79 


76.70 


68.10 


60.97 


56.10 


48.39 


71.90 


63.70 


56.82 


50.50 


45.04 


67.00 


59.20 


51.67 


46.70 


41.49 


62.30 


54.75 


48.59 


43.00 


38.83 


57.45 


50.25 


44.47 


40.10 


34.64 



600 

$85.00 

11.90 

4.5 

24.70 

5. 

1. 



40 

08 



43.08 
49.28 
46.18 
43.08 
3G.98 
36.88 
33.83 
30.73 



Size of plant horse power 

Cost of plant per horse power, 
Fixed charges at 14 per cent. . 
Coal per horse-power hour, pounds 
Cost of fuel at $4.00 per ton. . 
Attendance, 10-hour basis. . . . 

Oil, waste, supplies 

Total 
With coal at $5.00 per ton .... 
With coal at $4.50 per ton .... 

With coal at $4.00 per ton 

With coal at $3.50 per ton .... 

With coal at $3.00 per ton 

With coal at $2.50 per ton .... 
With coal at $2.00 per ton 



700 


800 


900 


1000 


1500 


$76.00 


$69.00 


$64.00 


$60.00 


$58.00 


10.65 


9.65 


8.95 


8.40 


8.12 


4.0 


3.5 


3.0 


2.5 


2.0 


22.00 


19.20 


16.50 


13.75 


11.00 


4.70 


4.15 


3.75 


3.50 


3.25 


0.94 


0.83 


0.75 


0.70 


0.65 


38.29 


33.83 


29.95 


26.35 


23.02 


43.79 


39.73 


34.05 


29.80 


25.77 


41.04 


36.28 


32.00 


28.05 


24.39 


38.29 


33.83 


29.95 


26.35 


23.02 


35.54 


31.48 


27.87 


24.60 


21.64 


32.79 


29.03 


25.80 


22.90 


20.27 


30.04 


27.18 


23.75 


21.20 


18.89 


27.29 


24.23 


21.70 


19.47 


17.52 



2000 

$56.00 

7.85 

1.5 

8.25 

3.00 

0.60 

19.70 

21.75 

20.72 

19.70 

18.67 

17.65 

16.60 

15.57 



712 



STEAM POWER PLANT ENGINEERING 



TABLE 101. 

SHOWING CAPITAL COSTS OF PLANTS INSTALLED AND ANNUAL COSTS OF POWER 
PER BRAKE HORSE POWER, AVERAGE WORKING CONDITIONS. 

(H.Von Schon, Engineering Magazine, May, 1907.) 

Class I. — Engines: Simple, Slide- Valve, Non-Condensing. 
Boilers: Return Tubular. 







Capital Cost of Plant per 


Annual Cost 


Annual Cost 


Size of Plant, 


Engines, 


Horse Power Installed. 


of 10 Hours 


of 20 Hours 


Horse Power. 


Boilers, etc., 
Installed. 




Power per 
B.H.P. 


Power per 








B.H.P. 






Building. 


Total. 






10 


$66.00 


$40.00 


$106.00 


$91.16 


$180.76 


20 


56.00 


37.00 


93.00 


76.31 


151.48 


30 


48.70 


35.00 


83.70 


66.46 


131.68 


40 


44.75 


33.50 


78.25 


59.49 


117.74 


50 


43.00 


31.00 


74.00 


53.95 


106.46 



Class II. — Engines: Simple, Corliss, Non-Condensing. 
Boilers: Return Tubular. 



30 
40 
50 
60 
80 
100 



$70.70 
62.85 
59.00 
56.00 
50.00 
44.60 



$35.00 
33.50 
31.00 
30.00 
27.50 
25.00 



$105.70 
96.35 
90.00 
86.70 
77.50 
69.60 



$61.14 
55.50 
50.70 
47.42 
43.86 
40.55 



$117.70 
107.10 
97.73 
91.34 
85.41 
79.19 



Class III. — Engines: Compound, Corliss, Condensing: 
Boilers: Return Tubular with Reserve Capacity. 



100 
150 
200 
300 
400 
500 
750 
1000 



$63.40 
53.70 
50.10 
45.90 
43.55 
41.25 
40.50 
39.00 



$28.00 
24.00 
20.00 
18.00 
16.00 
14.00 
13.00 
12.00 



$91.40 
77.70 
70.10 
63.90 
59.55 
55.25 
53.50 
51.00 



$33.18 
29.83 
28.14 
26.27 
24.84 
23.73 
23.56 
23.26 



$60.05 
54.63 
51.72 
48.83 
46.12 
44.21 
44.02 
43.71 



FINANCE AND ECONOMICS — COST OF POWER 



TABLE 102. 

COST OF ELECTRICAL POWER. (W. M. WILSON). 
(Power, January, 1907, p. 15.) 



0. < 


*1 

Hi o 

a. z 


CL< 

ii 


£cc 

J a. 


si 
Si 

(Old 


(/) 
UJ 

z 
o 
z 
u 

U. 

o 

6 

z 


8 u 


§1 


H 


8 

Hid 

§! 


2 
o 

"■ ° 

O C 
1- u 


CO 


OCC 

5:5 

HO 


V) 

tec 

Oui 

Z £D 


j -i 

f> CO 


o 2 

u CO 


°g 
a 


HO. 

O U 

U uj 

L. 


O) 

a. 
u_ u 

si 

O 5 
o o 

u 


U.CA 
Off 
Id 
t-V- 
W< 
O Id 
OI 




O 


SIMPLE 

non-ccnd, 

H.B, 


100 


30 


100 


2 


2880 


2300 


1000 




1292 


92 


F.T. 


2 


02 


1904 


2680 


280 




160 




COMPOUND 

NON-CCNB. 

M.6, 


120 


24 


100 


2 


4600 


2300 


1000 




1430 


74 


F.T. 


2 


74 


1840 


2510 


230 




130 




COMPOUND 
CONO. H. 8. 


120 


20 


100 


2 


4600 


2300 


1000 


1400 


1430 


66 


F.T. 


2 


66 


1500 


2414 


200 




125 




DE I AVAL 
TURBINE 


120 


18.6 


100 


2 


7400 


1000 


1200 


910 


62 


F.T. 


2 


62 


1425 


2360 


200 




120 




© 
o 

CM 


SIMPLE 

NON-COftO. 

H.8. 


100 


30 


100 


3 


4290 


3450 


1500 




1830 185 


F.T. 


3 


93 


2&50 


3850 


530 




235 




COMPOUND 

NON-COND. 

M.S. 


120 


24 


1C0 


3 


C300 


8450 


1500 




2040|l46 


F.T. 


3 


74 


2460 


3600 


440 




200 




COMPOUND 
CONO. M. 6. 


120 


20 


100 


3 


6900 


8450 


1500 


2060 


2040 


132 


F.T. 


3 


66 


2250 


3440 


400 




180 




OE LAVAL 
TURBINE 


120 


18.6 


100 


3 


11100 


1500 


1940 


1270 


12j 


F.T. 


2 


124 


2480 


2825 


390 




180 




o 

§ 

o 

o 


SIMPLE 

NON-COND. 

M. 8. 


100 


30 


200 


3 


7020 


6630 


3000 




2450 


370 


F.T 


4 


123 


4920 


5312 


1000 




400 




COMPOUND 

NON-COND. 

M.S. 


120 


24 


200 


3 


10620 


6630 


3000 




2698 1297 


F.T. 


4 


99 


3960 


5098 


840 




335 




COMPOUND 
COND. H. 8. 


120 


20 


200 


3 


10620 


6630 


3000 


3700 


2700 |264 


F.T. 


4 


88 


3670 


4970 


740 




312 




OE LAVAL 
TURBINE 


120 


17 


200 


3 


18960 


8000 


3250 


1620 


22c 


F.T. 


3 


113 


3390 


4000 


650 




275 




VERTICAL 
COND. L. 6. 


150 


13 


400 


2 


20400 


10600 


4000 


2000 


4300 


153 


W.T 


2 


153 


6100 


2300 


500 


665 


200 




HOR. 
COND.L.S. 


150 


13 


400 


2 


13600 


15400 


4000 


2600 


5440 


153 


W.T. 


2 


153 


6100 


2300 


500 


6C5 


200 




SIMPLE 

NON-C0N0. 
M.S. 


100 


30 


200 


4 


9360 


8840 


4000 




3190 


555 


FT. 


6 


111 


6660 


. 7070 


1460 




560 




COMPOUND 
NON-COND. 

M.8. 


120 


24 


220 


4 


141C0 


8840 


4000 




3512 


445 


F.T. 


5 


111 


5550 


6440 


1200 




4U0 




COMPOUND 
COND H.8. 


120 


20 


200 


4 


14160 


8840 


4000 


5050 


3512 


397 


F.T 


5 


99 


4950 


6296 


1070 




420 




OE LAVAL 
TUROI" 1 ? 


T20 


17 


2u0 


4 


25280 


4000 


4500 


2094 


339 


F.T. 


4 


113 


4520 


5240 


930 




380 




VERTICAL 
COND. L. 6. 


150 


13 


300 


3 


29550 


12600 


4500 


3600 


6225 


230 


W.l 


2 


230 


8460 


2530 


650 


1150 


280 




HOR. 
COND. 1. 8. 


150 


13 


300 


3 


19800 


19920 


4500 


3600 


7CC0 


i3& 


W.T 


2 


230 


8460 


2533 


650 


1150 


2$0 




PARbONS 
TORPINF 


150 


13 


COO 


2 


35010 


6000 


3600 


1470 


230 


W.T 


2 


230 


8460 


2530 


650 


1150 


280 




o 
o 

CM 


PIMPIE 

NON-COND. 

M.8. 


100 


30 


<u0 


4 


15200 


18000 


8000 




4030 


mc 


F.T. 


10 


122 


12300 


12776 


2500 




1060 




COMPOUND 

NON-COND. 

H.8. 


120 


24 


400 


4 


22280 


18000 


8000 




4540 


89C 


F.T. 


8 


127 


10160 


10340 


2120 




tiCO 




Compound 

CONO. H. 6. 


120 


20 


400 


4 


22280 


18000 


8000 


8100 


4540 


794 


F.T. 


8 


113 


9040 


10128 


1950 




780 




OE LAVAL 
TURBINE 


120 


16 6 


300 


5 


450)0 


7600 


7200 


2960 


660 


F.T. 


7 


110 


7700 


8880 


1700 




660 




VERTICAL 
CONO. L. 6, 


150 


13 


600 


3 


43200 


21300 


9000 


6100 


7132 


460 


W.T 


3 


230 


12690 


3660 


1220 


2300 


480 




HOR. 
COND. L. 6. 


150 


13 


600 


3 


28800 


28020 


9000 


6100 


8420 


4G0 


W.T 


3 


23u 


12690 


3660 


1220 


2300 


480 




PARSONS 
TURBINE 


150 


13 


1200 


2 


45000 


12000 


6100 


2800 


460 


W.T 


3 


230 


12C90 


3660 


1220 


2300 


400 




© 

o 
o 

CM 


VERTICAL 
CONO. 1. 8 


. 150 


13 


500 


5 


60000 


31350 


12500 


8450 


10735 


768 


W.T 


4 


250 


1C400 


4900 


1S20 


3840 


760 




HOR. 

COND.L.S 


150 


13 


1000 


3 


48000 


34650 


16000 


•8450 


9320 


76C 


WT 


4 


256 


18400 


4900 


1920 


3340 


700 




PAR60NS 
TURBINE 


150 


13 


1000 


3 


62400 


15000 


8450 


3360 


76£ 


W.T 


4 


256 


18400 


4£00 


1620 


3840 


700 





714 



STEAM POWER PLANT ENGINEERING 



TABLE 102 {Continued). 



U. 

O K 

CO K 




« CONTINUOUS EXPENSE PER YEAR 


J o 
< o 

H 

z 


INTEREST 
DEPRECIATION 
REPAIR? -TAX 
INSURANCE 


1* 

0$ 


HI 

O 

z 
< 

z 
Id 

t 

< 


(9 


COAL TOTAL 


UPPER LINE 10 Hrt. DAY 
LOWER LINE 24 HR. DAY 


is 
8 


COST OF COAL PER 2000 LB 


COST OF COAL PER 2000 LB 


$7 00 


$5.00 


$3.00 


$1.50 


$7.00 


$5.00 


$3.00 


11.50 


4G0 


12956 | 1814 


102 


1650 


4474 


3197 


1913 


959 


7940 


6663 


6384 


4426 


289 


4380 


11234 


8024 


4814 


2407 


17717 


14507 


11297 


8890 


.296 


14136 | 1979 


102 


1550 


3594 


2568 


1540 


770 


7225 


6198 


5171 


440I 


289 


4380- 


9022 


6444 


3867 


1933 


15670 


13092 


10514 


8581 


264 




102 


1550 


193 3206 


2291 


1374 


687 


7184 


6269 


6352 


4665 






289 


4380 


646 


8049 


5749 


3449 


1725 


15397 


13097 


10797 


9073 


248 


14863 


2081 


102 


1550 


180 


2986 


2133 


1280 


640 


6899 


6046 


5193 


4553 


289 


4380 


508 


7478 


6348 


3209 


1604 


14736 


12606 


10467 


8862 


740 


19276 


2699 


205 


2170 




8949 


6392 


3836 


1918 


14023 


11466 


8910 


6992 


678 


6132 




22468 


16048 


9628 


4814 


31877 


25457 


19037 


14223 


592 


21182 


2965 


205 


2170 




7188 


5135 


3081 


1540 


12528 


10475 


8421 


6880 


578 


6132 




18044 


12888 


7/33 


3867 


27719 


22563 


17408 


13542 


523 


22738 


3183 


205 


217U 


387 


6413 


4581 


2748 


1374 


.12358 


10526 


8693 


7319 


678 


6132 


1093 


16U97 


11498 


6899 


8449 


27083 


22484 


17885 


14435 


496 


22181 


3105 


205 


2170 


360 


5972 


4266 


2559 


128U 


11812 


IOI06 


8399 


7120 


578 


6132 


1016 


14956 


10696 


6418 


8209 


25787 


21527 


17249 


14040 


1480 




ifiin 


409~ 


3026 




17898 


12784 


7671 


8836 


25843 


20729 


15616 


11781 




1166 


8550 




44936 


32096 


19266 


9628 


59152 


46312 


83472 


23844 


1188 


34369 8 4812 


409 


3026 




14377 


10269 


6161 


3081 


22624 


16510 


14408 


11328 


1156 


8550 




86088 


25777 


15466 


7733 


50606 


40295 


29984 


22251 


1056 


37398 


5236 


409 


8026 


773 


12825 


9101 


6497 


2748 


22269 


18605 


14941 


12192 


1166 


855U 


2187 


32195 


22996 


18798 


6899 


49324 


40125 


80927 


24028 


904 


36049 


6047 


409 


8026 


658 


10913 


7795 


4677 


2339 


20053 


16935 


13817 


114/tf 


1156 


8650 


1858 


27373 


19552 


11731 


5866 


43984 


36163 


28342 


22477 


612 


52277 


7318 


409 


3026 


603 


7420 


5300 


8180 


1590 


18676 


16556 


14436 


12847 


TH56 


8550 


1421 


18615 


13296 


7978 


8988 


37060 


31741 


26423 


22433 


612 


51417 


7 198 


409 


3026 


603 


7420 


53U0 


31 au 


1590 


18556 


164.36 


U316 


12727 


1166 


855u 


1421 


18615 


13296 


7978 


3988 


36940 

37418 


29747 


26303 


22313- 


2220 


43960 


6153 


614 


3813 




26848 


19177 


11506 


5763 


22076 


16323 


1734 


10775 




67404 


48144 


28884 


14442 


86056 


66796 


47536 


33094 


'1780 


45942 


6432 , 


614 


3813 




21565 


15404 


92« 


4621 


32424 


26263 


20100 


15480 


1734 


10775 




54132 


88665 


23199 


llDUO 


73073 


57606 


42140 


80541 


1588 


49886 


6984 


614 


3813 


1161 


19238 


13733 


8245 


4122 


31810 


26305 


20817 


16694 


1734 


10775 


3279 48292 


34494 


20697 


10348 


71064 


67266 


.43469 


38120 


1356 


48300 


6762 


ew 


8813 


986 


16370 


11693 


7016 


8508 


, 28545 


2386b 


19191 


16C83 


1734 


10775 


2783 


41059 


29328 


17597 


8799 


63118 


61387 


8905b 


80853 


920 | 69465 Q79^ 


614 


8813 


755 


11130 


7950 


4770 


2386 


26037 


22857 


19677 


17293 






1734 


10775 


2132 


27923 


19945 


11967 


5983 


52289 


44311 


36833 


30849 


920 69490 


9729 


614 


3013 


755 


11130 


7950 


4770. 


2386 


26041 


22861 


19681 


17297 


1734 


10775 


2132 


27923 


19345 


11967 


5983 


44315 


86337 


30353 


920 60060 


8408 


614 


8813 


755 


1113U 


7950 


4770 


2386 


24720 


21540 


18360 


16976 


1734 


10776 


2132 


27923 
53694 


19945 

38353 


11067 


5983 


50972 


42994 


35016 


29032 


4440 


78306 


10963 


1227 


6580 




23012 


11506 


71464 


56123 


40782 


29276 


34b9 


16768 




134BU3 


96288 


67768 


28884 


165008 


126488 


87968 


69084 


3560 


79860 


11180 


YM 


6580 




43i2a 


30807 


184*0 


9242 


61116 


48794 


36472 


27229 


34b9 


15768 




108263 


77331 


• 46398 


23200 


138680 


107748 


76815 


63617 


3176 1 85S94 


12039 


1227 


6580 


232i 


38476 


274C5 


10488 


8246 


59643 


48632 


37656 


29412 






34o9 


15768 


6558 


965o4 


68989 


4l3aa 


20697 


134418 


106823 


79227 


68531 


2640 


84240 


11794 


1227 


6580 


1925 


319b5 


22830 


13699 


6850 


62491 


43366 


34225 


27376 


3409 


15768 


5440 


8020? 


57290 


34374 


17187 


116678 


93761 


70846 


63658 


1840 


108992 


16249 


1227 


6580 


1509 


22260 


15900 


9640 


4772 


45825 


39465 


3311)5 


28337 


34b0 


15768 


4264 


65846 


89890 


23934 


11969 


94596 


78640 


62684 


60719 


1840 


102530 


14354 


1227 


5580 


1509 


2226U 


15900 


9540 


4772 


44930 


36670 


32210 


27442 




3469 


16768 


4264 


5584G 


39890 


23934 


11969 


93701 


77745 


61789 


49824 


1840 1 88090 | 12333 


1227 


5580 


1509 


22260 


15900 


954U 


4772 


42909 


36549 


30189 


25421 


3469 


16768 


4264 


65846 


39890 


23934 


11969 91680 


75724 


59768 


47803 


3072 


155927] 21830 


2046 


7440 


' 2515 


37100 


26500 


15901 


7952 70931 


60331 


49732 


41783 


5782 


21025 


7106 


93076 


66483 


39890 


1994 148818 


122225 


95632 


76685 


3072 


146312 I 20764 


2046 


7440 


2515 3710U 


26500 


lOaU 


7952 


69865 


59265 


48666 


40717 


6782 


21025 


7106 


93076 


66483 


39890 


19943 


147752 


121159 


945G6* 


74619 


3072 


122102! 17094 


2U4S 


7440 


2515 


871UU 


26500 


15901 


7952 66195 


55595 


44996 


37047 




"' y 


6782 


21025 


7108 93076 


66483 


39890 


19943 y 144082 


117489 


90896 


70949 ] 



FINANCE AND ECONOMICS — COST OF POWER 



715 



TABLE 103. 

COST OF ONE HORSE POWER PER YEAR IN STREET-RAILROAD SERVICE FOR 
DIFFERENT CLASSES OF ENGINES. 



1000-Horse-Power Plant. (R. C. Carpenter.) 
(Sibley, Journal of Engineering, November, 1904, p. 92.) 



Non-Condensing Engines. 

Simple slide-valve, aver- 
age 

Simple slide-valve, best. . . 

Simple Corliss, average 

Simple Corliss, best 

*Compound slide-valve. . . 
Condensing Engines. 

Compound slide-valve 
average 

Compound slide-valve, 
best 

Compound Corliss, average 

Compound Corliss, best. . . 



CD 

si 

1 & 

a 


Tons per Horse 
Power per Year. 


Fuel Cost per Year, 18 
Hours per Day, $2.00 
per Ton. 


o 

-8 

Hi 

a 
»" 

6 


Interest, 5 per Cent; 
Depreciation, 10 per 
.Cent. 


O 

M 

7 

Q 


S 

00 

7 

Q 


O 

M 

1 

S3 1 




4.63 


10.14 


15.21 


20.28 


30.42 


7.30 


4.44 


4.60 


10.07 


15.10 


20.14 


30.20 


7.30 


4.44 


3.45 


7.55 


11.33 


15.10 


22.66 


7.30 


4.75 


3.01 


6.59 


9.89 


13.18 


19.78 


7.30 


4.75 


4.17 


9.05 


13.57 


18.10 


27.14 


6.90 


4.80 


3.25 


7.12 


10.68 


14.24 


21.36 


6.70 


4.72 


2.40 


5.25 


7.88 


10.51 


15.76 


6.50 


4.72 


2.36 


5.17 


7.74 


10.33 


15.48 


6.50 


5.28 


1.80 


3.94 


5.91 


7.88 


11.82 


6.10 


5.28 



c3 o 
O W 



42.16 
41.94 
34.71 
31.83 
38.84 



32.78 

26.98 
27.26 
23.20 



* The compound slide-valve engine, running non-condensing, made, in this series of tests, a 

poorer record than the single Corliss. This may have been due to the extremely bad conditions 

of loading. It is, I think, a fact that this class of engine has not been a marked success for 
street railway work. 



716 



STEAM POWER PLANT ENGINEERING 



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sasuadxg jinox 



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jnojj av^; jad 
sasuadxg; jb^ojl 



s^uaQ — jajj^g jad 

s^uao Off pnj t^im 

jnojj ayjj jad 

sasuadxg; pjox 



s;uao — lajj-eg 

jadOO'TS^- 1110 !! 
■ai^j jad janj 



s^uaQ — jnojj -Ai^j 
jad aa^-BAi P U,B 
'saijddng 'joq-ei 



I!0 3° P"^a 
jad sanojj "aijj; 



■s^ua^ — jnojj 
•Aijj jad jaiB^ 



•s^ua^ — jnojj 
Ai^j jad saijddng 



•^sa X Jno H -2T HO P 
jajj^g jad sjnojj -ayjj 



sjbjjoq— j'eaj^ jad 
joq^ jo ^soq \viojl 



•sj^ijoq — J'eax J 9 d 
saSj^qo paxij jt^ox 



' s ii B AV Su ?pn n a p a P! s 

-;nQ Sntdjj jo a^sg 

IBa-jj aprip'uj !jou saoQ 

•sj^pa— intJfj 

jo ^soq isjij p^6x 



■mj — s^iuq jcazig 



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TABLE 105. 

COST OF POWER. 
Examples of Isolated Station Practice. 



Rated capacity kilowatts 

Yearly capacity kilowatt hours 

Actual load kilowatt hours 

Yearly load factor per cent 

Curve load factor per cent 



Large Office 
Building. 



500 

4,380,000 

670,000 

17 

17 



Small Office 
Building. 



50 
438,000 
40,820 
9.1 

24.7 



Manufactur- 
ing Plant, 
Electro- 
plating. 



500 

4,380,000 

3,500,000 

80 

80 



Operating Charges, per Year. 



Labor 

Coal and ashes 


$6,050.00 

6,342.00 

642.00 

168.00 

395.00 

69.00 


$1,400.00 

960.00 

75.00 

90.00 

41.00 

182.00 


$12,300.00 
9,100.00 


Water 


Oil and waste 


210 00 


Lamps 

Repairs and renewals 


50.00 
1,008.00 


Total 


$13,666.00 


$2,748.00 


$22,668.00 



Fixed Charges, per Year. 



Interest (5 per cent) 


$3,500.00 
4,200.00 

350.00 
1,050.00 

900.00 


$325.00 

628.00 

30.00 

90.00 


$4,500.00 


Depreciation (6 per cent) 


5,400.00 

450.00 

1,350.00 


Insurance (£ per cent) 

Taxes (11 per cent) 










Total 


$10,000.00 


$1,073.00 


$11,700.00 







Cost per Kilowatt Hour, Cents. 



Operating charges 
Fixed charges 
Total cost 




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FINANCE AND ECONOMICS — COST OF POWER 



723 



TABLE 107a. 

YEARLY OPERATING COSTS IN FOUR TYPICAL CENTRAL STATIONS, 

STATE OF MASSACHUSETTS 

Year ending June, 1909. 



Type of Prime Mover 

Rated station capacity, kw.. 
Output, millions of kw. hrs.. . 
Yearly load factor, per cent. . 
Total station operating force 
Cost of fuel, dollars per ton . . 
Coal per kw. hr 



Salem Elec- 
tric Light 
Co. 



6 Engines 

2500 

3.106 
14.2 
14 

4.51 

3.3 



Fitchburg Gas 

& Electric 

Co. 



3 Engines 

2000 

4.006 
22.9 
12 

4.52 

3.28 



Haverhill 
Electric Co. 



2 Turbines 
1 Engine 

2300 

3.721 
18.5 
13 

3.97 

3.27 



Maiden 
Electric Co. 



1 Turbine 
3 Engines 



4.715 



14 

3.78 
3.02 



Operating 


Costs, Cents 


per Kilowatt Hour. 




Coal 


0.740 
0.025 
0.027 
0.410 
0.034 
0.158 
0.011 


0.740 
0.015 
0.025 
0.308 
0.017 
0.041 
0.072 
0.024 


0.650 
0.190 
0.003 
0.285 
0.063 
0.073 
0.019 
0.040 


0.565 


Oil and waste 


0.020 


Water 


0.045 


Wages 


0.320 


Station building repairs. .... 
Steam equipment repairs. . . . 
Electrical equipment repairs. 
Miscellaneous 


0.023 
0.072 
0.14 
0.21 








Total 


1.412 


1.242 


1.152 


1.08 



TABLE 107b. 
COST OF POWER, CENTS PER KW. HOUR. STEAM-ELECTRIC CENTRAL STATIONS. 

Year ending June 30, 1908. 



Fuel 

Oil and waste 

Water 

Wages 

Station repairs 

Steam repairs 

Electrical repairs 

Miscellaneous 

Total 

Cost of fuel per ton 

Output, millions kilowatt 
hours per year 

Capacity of station, thou- 
sands of H.P 



Bos- 
ton. 



.462 
.008 
.024 
.192 
.015 
.042 
.056 
.023 



.822 
3.99 

88.5 

73.5 



Worcester. 



.703 
.027 
.034 
,360 
,012 
055 
055 
000 



1.246 
4.79 

5.4 

5.90 



Lowell 



.710 
.009 
.008 
.262 
,020 
,020 
,009 
022 



1.060 
4.75 

9.4 

7.39 



Fall 
River. 



.880 
.032 
.012 
.538 
.012 
.037 
.029 
,080 



620 
68 



4.0 
4.43 



Mai- 
den. 



.635 
.017 
.032 
.342 
.035 
.072 
.014 
.033 



1.180 
4.49 

4.6 

4.87 



Cam- 
bridge. 



.690 
.019 
,055 
,347 
,021 
059 
046 
000 



1.237 
4.40 

6.0 

6.75 



Lynn. 



.618 
.012 
.040 
.296 
.052 
.147 
.045 
.000 



1.210 
3.60 

8.7 

8.2 



724 



STEAM POWER PLANT ENGINEERING 



TABLE 108. 

COST OF POWER (1907), FIRST NATIONAL BANK BUILDING, CHICAGO. 



Total weight of coal burned tons 

Total weight of water evaporated gallons 

Total electrical output kilowatt hours 

Water actually evaporated per pound of coal as fired 



14,956 
22,100,000 
1,546,600 
6.1 



Electric-Light Plant. (Cost of Power only.) 



Pounds of coal per kilowatt hour 

Cost of coal per kilowatt hour cents 

Cost of labor per kilowatt hour cents 

Cost of supplies per kilowatt hour cents 

Total '. 



6.57 



0.88 
0.78 
0.11 

1.77 



All expenses of entire plant charged against switchboard. 



Pounds of coal per kilowatt hour 

Cost of coal per kilowatt hour cents 

Cost of labor per kilowatt hour cents 

Cost of supplies per kilowatt hour cents 

Total 



19.34 



2 


23 


1 


59 


1 


18 



5.00 



Elevator Plant. 



Passengers handled 

Total car miles 

Cost of labor per car mile cents 

Cost of material per car mile cents 

Cost of power * per car mile cents 

Total 



2,016,300 
92,700 



8.14 



* Approximate. 

BIBLIOGRAPHY. COST OF POWER. * 
Electrical. 

Rented Power for Electric Railways, American 

Electrician 10: 329 July, 1898 

Cost of Pumping Station in New York, Electrical 

World 43: 820 April 30, 1904 

Cost of Electric Power in Comparison with Steam 

for Traction, Engineer (London) 90: 600 Dec. 14, 1900 

Cost of Steam and Electricity, Engineering 74: 667 Nov. 21, 1902 

Cost of Power, Engineering 76: 706 Nov. 20, 1903 

Economy of Isolated Plant (I. D. Parsons), 

Engineering Magazine 22: 573 Jan., 1902 

720 Feb., 1902 



* See page 729 for bibliography 1908-1910. 



FINANCE AND ECONOMICS — COST OF POWER 



725 



Electrical — Continued. 
Cost of Energy in Electric Supply (A. D. Adams), 

Engineering Magazine 24 : 

Street Railway Review 12: 

Data on Electric Power Generation in Glasgow, 

Engineering Record 

Cost of Generating Electric Power (E. J. Fox), 

Engineering Record 49 : 

Relative Cost of Electric Power for Three Types of 

Plants (R. D. Mushon), Engineering Record 49: 

Cost of Electric Power for Street Railways (R. W. 

Conant), Power 18 : 

Street Railway Review 8: 

Cost of Power in New Orleans Railroad Company 

Power Plant, Street Railway Journal 18: 

Cost of Power, Union Traction Company of 

Indiana, Street Railway Journal 18: 

Relative Costs of Steam and Polyphase Traction, 

Street Railway Journal 21 : 

Cost of Power at Newcastle-on-Tyne, Street Rail- 
way Journal 22 : 

Charges for Rented Power, Street Railway Review. . . 8 : 
Cost of Power for Electric Railways, Street Railway 

Review 8 : 

9: 



181 


Nov., 1902 


149 


March, 1902 


478 


April 22, 1899 


388 


March 26, 1904 


411 


April 2, 1904 


8 


Oct., 1898 


631 


Sept., 1898 


668 


Nov. 2, 1901 


827 


Dec. 7, 1901 


737 


May 16, 1903 


207 


Aug. 8, 1903 


236 


April, 1898 



43, 97, 186, 224, 340, 385, 
461, 760, 886, 1898 
35, 123, 185, 261, 319, 
459, 529, 595, 749, 851 
10: 11, 223, 399, 521, 735 
11:123,416, 418, 1901 
Cost of Niagara Power at Buffalo, Street Railway 

Review 8: 339 May, 1898 

Statistics on the Cost of Power, Street Railway 

Review 12: 77 Oct., 1902 

Analysis of Cost of Generation and Distribution 
of a Unit of Electricity (C. W. Rice), Western 

Electrician 22: 574 June 25, 1898 

Cost of Electric Power at Lachine Rapids, Canada 

(W. L. Walbank), Western Electrician 23: 24 July 9, 1898 

Suggestions Relative to Determining Cost of Elec- 
tric Supply (M. E. Turner), Western Electrician. . . 23: 143 Sept. 10, 1898 
Cost of Power (C. S. Brown), Western Electrician ... 28: 127 Feb. 23, 1901 

Cost of Power (C. Grey), Western Electrician 30: 211 March, 1902 

Cost of Power (M. J. Eichorn), Western Electrician. . 31 : 69 Aug. 2, 1902 

Graded Costs of Electrical Supply (M. E. Turner), 

Western Electrician 35: 204 Sept. 10, 1904 

Some Notes on the Cost of Generating Electrical 

Energy (E. J. Fox), Engineer (London) 27: 219 Feb. 26, 1904 

269 March 11, 1904 
Effect of Load on the Cost of Power (E. M. Archi- 
bald), Engineer (United States) 42: 315 May 1, 1905 

Cost of Power in Street Railway Service, Machinery 11:317 Feb., 1905 



726 STEAM POWER PLANT ENGINEERING 

> 

Electrical — Continued. 

Cost of Electric Power at the Switchboard (C. H. 

Hile), Power 25: 662 Nov., 1905 

Power Plant Economies (H. G. Stott), Engineer 

(United States) 43: 191 March 1, 1906 

Power Costs (Charles E. Lucke), Electrical Review 

(New York) 50: 797 May 18, 1907 

Systems of Charging for Electricity Supply (W. A. 

Toppin), Electrical Engineer (London) 39: 42 Jan. 11, 1907 

Rates of Charge for Electricity and Their Effect 

on Cost (J. S. Codman), Proceedings of the 

American Institute of Electrical Engineers 26: 31 April, 1907 

The Principles of Modern Rate-Making for Electric 

Light and Power, Electrical World 49: 1086 June 1, 1907 

Methods of Computing Central Station Rates in 

Boston, Electrical World 49: 1090 June 1, 1907 

The Present Tendency of Charging for Electricity 

(W. A. Toppin), Electrical Review (London) 60: 945 June 7, 1907 

Electric Power Tariffs (C. S. Nesey-Brown), 

Cassier's Magazine 32: 304 Aug., 1907 

The Sale of Electricity for Lighting Purposes (L. E. 

Bucknell), Electrical Engineer (London) 40: 370 Sept. 13, 1907 

Rates and Systems of Charging (Jacques), Elec- 
trical Review (London) 61: 1074 Dec. 27, 1907 

Gas. 

Cost of Pumping Station in New York, .Electrical 

World 43: 820 April 30, 1904 

Comparative Cost of Power Generated by Steam 

Engine, Water Turbine, and Gas Engine, Engineer 

(London) 88: 320 Sept. 29, 1899 

Comparative Cost of Generating Power by Steam 

Engine, Water Turbine, and Gas Engine (J. B. C. 

Kershaw), Engineering 70: 351 Sept. 14, 1900 

390 Sept. 21, 1900 

Cost of Gas Power for Central Station, Engineer .... 71 : 27 Jan. 4, 1901 

Gas Power for Central Stations (J. R. Bibbins), 

Engineering Record 49: 11 Jan. 2, 1904 

Street Railway Journal 1089 Dec. 26, 1903 

Power 24: 100 Feb., 1904 

Comparative Cost of Steam and Gas Plant, 

Engineering Record 49: 310 March 5, 1904 

Is Gas Power More Economical than Water 

Power ? (H. C. T. Horace), Power 25: 599 Oct., 1905 

Cost of Steam Power (Edit), American Electrician . . 10: 114 March, 1898 

Cost of Steam Electrical Generating Plant (R. C. 

Carpenter), Electrical World 43: 1016 May 28, 1904 

Economy of Power Installations (C. Weiss), 

Engineering 66: 59 July 8, 1898 

Estimates for an Electric Light Plant in New York 

City, Engineering News 52: 583 Dec! 29, 1904 



FINANCE AND ECONOMICS — COST OF POWER 



727 



Gas — Continued. 

Cost of a Power Station in Europe, Street Railway 

Journal 20:210 Aug. 9, 1902 

Power House Cost, Louisville Electric Railway, 

Street Railway Review 9: 592 Sept., 1899 

Miscellaneous. 

Improvements in Economy of the Steam Engine 

(W. F. Durand), American Electrician 11: 13 Jan., 1899 

68 Feb., 1899 

Cost of Power (Edit), Engineer (London) 96: 285 Sept. 4, 1901 

Estimating the Cost of Power, Engineer (United 

States) 36:285 Dec. 1, 1899 

Investigation of the Cost of Power (C. G. Gray), 

Engineer (United States) 39:43 Jan. 1, 1902 

Efficient Use of Steam and Labor in Isolated 

Plants (P. R. Moses), Engineering Magazine. ... 16: 99 Oct., 1898 

Cost Determination in Isolated Plants (P. R. 

Moses), Engineering Magazine 20: 1082 March, 1901 

Cost of Pumping at a Colliery (R. V. Norris), 

Engineering News 49: 228 March 12, 1903 

Economic Power Production (R. H. Thurston), 

Engineering Record 47: 35 Jan. 3, 1903 

Cost of Power (Carpenter), Power 24: 425 July, 1904 

Decreasing Costs in the Steam Plant (Edit), 

Engineer (United States) 42:412 June 15, 1905 

The Economy of a Small-Sized Coal for the Power 

Plant (P. R. Moses), Engineering Magazine 28: 783 Feb., 1905 

Analysis of Central Station Costs and Revenues 

(H. S. Knowlton), Engineering Magazine 29: 238 May, 1905 

Cost of Operating Buildings, Engineering Record 48: 759 Dec. 19, 1903 

Power Plant Supervision and Accounting (F. W. 

Ballard), Engineering Record 51 : 687 June 17, 1905 

Power Plant Economy, Power 25: 602 Oct., 1905 

Relative Efficiency and Desirability of Various 

Types of Engines (A. W. Richter), Street Railway 

Review 10: 162 March 15, 1900 



Steam. 

Comparative Cost of Power Generated by Steam 
Engine, Water Turbine, and Gas Engine, 
Engineer (London) 88: 322 

Cost of Electric Power in Comparison with Steam 

for Traction, Engineer (London) 90 : 600 

Investigation of the Cost of Power, Engineer 

(United States) 39: 161 

Cost of Steam Power per Horse Power Year 

(J. M. Whitman), Engineer (United States) 40: 741 

Cost of Steam Raising (J. Holliday), Engineering ... 68: 739 



Sept. 29, 1899 

Dec. 14, 1900 

March 1, 1902 

Oct. 1, 1903 
Dec. 8, 1899 
Dec. 22, 1899 



T28 STEAM POWER PLANT ENGINEERING 

Steam — Continued. 

Comparative Cost of Generating Power by Steam 

Engine, Water Turbine, and Gas Engine (J. B. C. 

Kershaw), Engineering 70: 351 Sept. 14, 1900 

390 Sept. 21, 1900 

Cost of Steam and Electricity, Engineering 74: 667 Nov. 21, 1902 

Comparative Cost of Steam and Water Power 

(W. O. Webber), Engineering Magazine 15: 923 Sept., 1898 

Steam Costs in an Industrial Combination (W. D. 

Ennis), Engineering Magazine 28: 86 Oct., 1904 

Fuel Economy of Engines in Electric Railway 

Power Stations (Carpenter), Engineering News. . . 42:234 Oct. 12, 1899 

Reduction in Cost of Steam Power from 1870 to 1897, 

Engineering Record 37: 12 Dec. 4, 1897 

Economy in Use of Coal for Production of Power 

(I. N. Hollis), Engineering Record 46: 491 Nov. 22, 1902 

Economy of Fuel in Electric Plants (Edit), 

Engineering Record 48: 233 Aug. 29, 1903 

Economical Steam Making (Edit), Engineering 

Record 48: 385 Oct. 3, 1903 

Cost of Fuel and Power in the South, Power 18:13 Nov., 1898 

Economical Production of Steam with Special 

Reference to the Use of Cheap Fuel, Power 19: 19 June, 1899 

Suggestions for Steam Economy (W. M. Kay), 

Engineer (United States) 42: 655 Oct. 2, 1905 

Yearly Cost of One Steam Horse Power, Machinery . 9 : 374 March, 1903 

Water. 
Cost of Water and Electric Power (G. E. Walsh), 

American Electrician 16: 331 July, 1904 

Comparative Cost of Power Generated by Steam 

Engine, Water Turbine, and Gas Engine, Engineer 

(London) 88: 322 Sept. 29, 1898 

Investigation of the Cost of Power, Engineer 

(United States) 39: 64 Jan. 15, 1902 

Comparative Cost of Generating Power by Steam 

Engine, Water Turbine, and Gas Engine (J. B. C. 

Kershaw), Engineering 70: 351 Sept. 14, 1900 

390 Sept. 21, 1900 

Cost of Water Power in France, Engineering 76: 571 Oct. 23, 1903 

Comparative Cost of Steam and Water Power 

(W. O. Webber), Engineering Magazine 15: 923 Sept., 1898 

Cost of Hydraulic Transmission of Power (E. B. 

Ellington), Engineering Magazine 17: 233 May, 1899 

Cost of Hydraulic Power in Switzerland, Engineer- 
ing Record 41 : 182 Feb. 24, 1900 

Analysis of the Commerical Value of Water Power 

(A. F. Nagle), Engineering Record 46: 540 Dec. 6, 1902 

Costs of Pumping Water, Power 20:12 Nov., 1900 



FINANCE AND ECONOMICS — COST OF POWER 



729 



COST OF POWER. 

(1908-1910.) 

Approximate Cost of Gas Power (M. P. Cleghorn), 

Power and Engineer April 7, 1908 

Central Station vs. Private Plants, Engineering Feb. 26, 1909 

Comparative Cost of Power Production, Electrical Age 40:63 March, 1909 

Electrical World 53: 792 April 1, 1909 

Electrical Review and Western Electrician 55: 773 April 28, 1909 

Cost of a Gas Engine and of a Combined Steam 

1 Plant, Engineering Record 60: 272 Sept. 4, 1909 

Cost of a Kilowatt-Hour (R. A. Day), Electrical World 54 : 853 Oct. 7, 1909 

Cost of Power (H. G. Stott), Pro. Am. Inst. Elec. 

Engrs 28: 283 April, 1909 

Cost of Power in a 3000 Kw. Turbine Plant, Elec- 
trical Review and Western Electrician 55: 62 Oct. 2, 1909 

Cost of Power in a 1500 Kw. Central Station, Engi- 
neering News : 61 : 471 April 29„ 1909 

Cost of Power in Small Plants (W. E. Snow), Engi- 
neering Magazine 33: 169 May, 1908 

Cost of Power in Four Central Stations, Electrical 

World 55: 813 March 31, 1910 

Cost of Power for Various Industries (C. T. Main), 

Jour. Assoc. Eng. Soc 44: 151 March, 1910 

Engineering Record 60: 711 Dec. 25, 1909 

Cost of Power in Varying Units (W. O. Webber), 

Engineering Magazine 35: 562 July, 1908 

Cost Systems and Time Keeping, Columbus, O., Muni- 
cipal Electric Lighting Plant, Engineering News Dec. 3, 1908 

Electric Power Costs in Small Stations, Engineering 

Record 59: 30 Jan. 9, 1909 

First Cost of Plant and Cost of Generating and 
Distributing Electricity for Lights, Brooklyn Edi- 
son Co., Engineering Contractor 33: 393 April 6, 1910 

Isolated Power Plant Costs and Their Relation to 

Central Station Service(W.F.~Lyod),ElectricalWorld 53:323 Feb. 4, 1909 

Isolated Station Records and Accounting Power .... April 28, 1908 

Operating Costs of Large Units, Power and Engi- 
neer 31: 981 May 31, 1910 

Power Costs for Factories, Engineering Record 60: 604 Nov. 27, 1909 

Power Plant Waste (P. R. Moses), Cassier's Maga- { 36: 497 Oct., 1909 

zine . . . ( 37: 12 Nov., 1909 

Relation of Load Factor to Power Costs, Jour. Wes. 

Soc. Engr 14: 241 April, 1909 

Engineering Record 59: 702 June 5, 1909 

Representative Data from Electric Power Plant 

Operation (H. S. Knowlton), Engineering Magazine 36 : 833 Feb., 1909 

Systems of Charging for Electrical Energy (W. T. 

Ryan), Engineering Magazine April, 1909 

The Small Station and its Economical Operation, 

Western Electrician 43: 10 July 4, 1908 

The Valuation of Steam Power Plants (C. T. Main), 

Electrical Age 39: 228 Oct., 1908 

Useful Figures from Practical Power Plant Opera- 
tion, Electrical World 54: 781 Sept. 30, 1909 

Working Results from a Gas-Electric Power Plant 

(J. R. Bibbins), Pro. Am. Inst. Elec. Engrs 27: 1123 July 1, 1909 



CHAPTER XVIII. 

TESTING AND MEASURING APPARATUS. 

403. General. — The importance of maintaining a system of records 
has been referred to in paragraph 388. The various items which may 
be recorded and the instruments and appliances used in this con- 
nection are outlined in the accompanying chart. In large stations a 
full complement of indicating, recording, and integrating instruments 
may prove to be a good investment if intelligently and closely studied 
by the operating engineer with a view to locating and eliminating 
unnecessary losses. The instruments should be inspected and cali- 
brated at intervals, since many of them are delicately constructed 
and are apt to become inaccurate after a few months' service. Steam 
gauges, thermometers, and pyrometers, and particularly water meters 
are subject to appreciable error after considerable use. Voltmeters, 
ammeters, and other switchboard instruments are easily deranged, espe- 
cially when subjected to continuous vibration or to high temperature. 

404. Weighing the Fuel. — In most small plants the delivery 
tickets of the coal dealer are depended upon for the weight of coal 
used, no attempt being made to determine its evaporative value, and 
the economy of the plant is judged by the size of the coal bill. In 
such cases a considerable saving can be effected by keeping a daily 
record covering at least the coal and water consumption. The coal 
can be conveniently weighed on ordinary platform scales. In a num- 
ber of large stations the weight of coal is determined by suspended 
weighing hoppers, which may be stationary, as in Fig. 109, or mounted 
on a traveling truck, as in Fig. 110. The scales of such devices are 
made indicating, autographic, integrating, or a combination of the 
three, the latter costing but little more than the simple indicating or 
recording devices. 

405. Measurement of Water. — The most accurate means of measur- 
ing water is by the use of two or more tanks resting upon scales, 
arranged to be filled and emptied alternately. This method, however, 
involves considerably more time than is ordinarily at the disposal 
of the fireman or engineer. The usual practice is to place a hot-water 
meter on the pressure side of the feed pump, with provision for 

calibration without shutting off the feed supply to the boilers. 

730 



TESTING AND MEASURING APPARATUS 



731 



TESTING AND MEASURING APPARATUS 

STEAM PLANT. 



Weights , 



Pressures. 



Tempera- 
tures 



Power. 



Fuel. 



Water .(Water meters 



Steam 



High 



Low. 



Flue gas 
analysis 



Moisture. 



Fuel analysis 



Platform scales, indicating and autographic. 

Suspension hoppers, indicating and auto- 
graphic. 

Platform scales and tanks. 
Piston. .. .") 
Rotary . . . ^Integrating. 

Disk J 

Venturi, indicating and 
autographic. 

Weirs. 

'Weighing condensed steam. 

Steam meters I St * John,s > autographic, 
steam meters. • | Burnham ^ indicating. 

( Bourdon gauge, indicating and autographic, 
t Manometers, mercurial, indicating. 
''Manometers — mercurial, indicating, and 

autographic. 
Manometers — water, indicating, and auto- 
graphic. 
.Diaphragms, indicating and autographic, 
f Mercurial thermometers, indicating. 
Up to 800 deg. F. 4 Expansion thermometers, indicating and 
[ autographic. 
Expansion thermometers, indicating and 

autographic. 
Resistance thermometers, indicating and 

autographic. 
Thermo-electric thermometers, indicating 

and autographic. 
'Optical pyrometer, indicating and auto- 
graphic. 
^Platinum or clay ball pyrometer. 

/T _j.-- „ + _j (Indicators, hand manipulated. 

f inaicaiea j Indicators, continuous autographic. 

{Rope brake. 
Prony brake 
Absorption dynamometers. 
Electric generator. 
TOrsat apparatus. 
J Arndt's econometer, indicating. 
] Sarco and Ados recorder, autographic. 
LUehling gas composimeter, autographic. 

fin air Hygrometer, indicating and autographic. 

1 1n steam Calorimeters. . - (Separating. 



■ 800 to 2500 deg. F. < 



Over 2500 deg. F. - 



Coal calorimeters 



Gas calorimeter. 



"• /Throttling. 

fMahler bomb. 
J Carpenter. 
1 Thompson. 

IParr. 
. .Junker. 



ELECTRICAL PLANT. 

Voltage Voltmeters, A.C. and D.C., indicating and autographic. 

Current Ammeters, A.C. and D.C., indicating and autographic. 

Output Wattmeters, A.C. and D.C., integrating and autographic. 

Power factor .Power factor meters, A.C. only, indicating and autographic. 

Frequency Frequency meter, A.C. only, indicating. 

Synchronism . .Synchronizers, A.C. only, indicating. 



732 



STEAM POWER PLANT ENGINEERING 



There are several types of meters in common use. Fig. 419 illus- 
trates the piston type, in which reciprocating pistons are displaced by a 
definite volume of water; Fig. 305, the rotary type, depending upon 




Fig. 419. A Typical Piston Water Meter. 

the displacement of rotating impellers; Fig. 420, the disk type, in 
which impellers are given a combined rotating and tilting motion. 
When periodically calibrated, water meters give satisfactory results. 

When graduated to read 
in pounds the accuracy de- 
pends upon the temper- 
ature range of the water; 
thus the density of water 
at 62 degrees F. is 62.36 
pounds per cubic foot and 
at 212 degrees it is 59.76, 
a range of 2.6 pounds per 
cubic foot. Hence a meter 
calibrated to read correctly 
at 62 degrees F. will have 
an error of about 4.2 per 
cent if used to measure 
water at 212 degrees. The 
average range in feed tern- 

Fxg.420. A Typical Disk Water Meter. perature fc se l d om greater 

than 40 degrees, and if the meter is calibrated for the mean tem- 
perature the error is somewhat less than one per cent. 

The Venturi meter, Fig. 421, is frequently employed for measuring 
large volumes of water, as in city waterworks, and in connection with 
condensing plants.* It amounts practically to a constriction in the 
diameter of the pipe, is readily installed, and the total absence of 

* Tests on a Venturi Meter for Boiler Feed. Proc. A.S.M.E., Mid.-Oct ., 1909, p. 1065. 




TESTING AND MEASURING APPARATUS 



733 



working parts is a great advantage. The meter is supplied with 
either indicating or autographic manometer. With water at constant 
temperature the error in the readings should not exceed one per 
cent. 

The pitometer is a simple adaptation of the well-known pitot tube, 
and is used for measuring the flow of water through pipes where it is 



IPES TO MANOMETER 



=U=4=i 



Fig. 421. Principles of the Venturi Meter. 



n^?n 



MANOMETER 



impracticable to insert a meter. It is only necessary to drill a small 
hole in the pipe for the introduction of the tube. The volume flowing 
may be calculated from the readings of the manometer or may be 
autographically recorded. 

In measuring large volumes of water flowing in open channels the 
measurements are made by weirs of suitable proportions or by current 
meters. 

Water Measurement, General: Eng. Rec, Feb. 15, 1902. 

Water Meters: Trans. A.S.M.E., 18-134, 14-676, 5-63; Engng. News, Jan. 3, 
1907, March 9, 1905, June 16, 1904, p. 569; Eng. Rec, Nov., 1903; Stevens Ind., 
Jan., 1901; Jour. New Eng. Waterworks Assn., June, 1907; Eleen., Lond. May 
8, 1908. 

Venturi Meter: Revue Technique, Feb. 10, 1905, Eng. News, Feb. 28, 1901; 
Prac. Engr., Feb. 15, 1907; Pro. A.S.M.E., Nov., 1906; Trans. A.S.C.E., Nov., 
1907, 57-531 (1906) ; Engng., Feb. 22, 1907, p. 236. 

Pitometer: Jour. Franklin Inst., Dec, 1907, p. 425; Tech. Quar., June, 1907; 
Jour. New Eng. Waterworks Assn., June, 1906; Trans. A.S.M.E., 25-184; Sib. Jour. 
Engng., May, 1902; Jour. Assn. Eng. Soc, Aug., 1901; Eng. News, March 31, 1904, 
Dec. 21, 1905; Pro. Engrs. Soc. of West. Penn., Dec, 1906. 

Pitot Tube: Am. Mach., Aug. 9, 1906, p. 175; Trans. A.S.C.E., 47-6, 57-265, 
25-184; Eng. News, March 31, 1904, p. 318, Dec 21, 1905, p. 660; Progressive Age, 
June, 1906, p. 63; Jour. Assn. Eng. Soc, Aug., 1901, p. 35; Sib. Jour., May, 1902; 
Cal. Jour, of Tech., May, 1905. 

Weirs: Engr. (Lond.), June 5, 1903, p. 562, Aug. 17, 1906; Am. Soc. Civ. Engrs., 
44-160; Eng. Rec, July 13, 1901, p. 32. 



>OWER PLANT ENGINEERING 

Current Meters: Cal. Jour, of Tech., April, 1904; Eng. News, March 7, 1907, 
p. 263, Feb. 12, 1902; Eng. Rec, Dec. 19, 1903; Pro. Am. Soc. Civ. Engrs., Sept., 
1901, Nov., 1901, Dec, 1901. 

Piezometers: Power, Aug., 1907, p. 569; Pro. Am. Soc. Civ. Engrs., 44-34, 49-112, 
51-252; Eng. News, Sept. 13, 1906, p. 271; Power, Aug., 1907, p. 569. 

406. Measurement of Steam. — In surface-condensing plants the 
weight of steam consumed by the engines is conveniently obtained 
by weighing the condensed steam; in practice, however, this method 




Fig. 422. St. John's Steam Meter. 



Fig. 422a. 



Principles of the Burnham 
Steam Meter. 



is adopted only when testing the plant, the feed-water measurement 
sufficing for general recording purposes. 

When steam is supplied to various points and the weights cannot be 
readily determined by condensing, steam meters are sometimes used. 
The St. John's, Fig. 422, is the best known in this country. This 
apparatus records the weight of steam passing through the seat of an 



TESTING AND MEASURING APPARATUS 



735 



automatically lifting valve which rises and falls as the demand for 
steam increases or diminishes. When the maximum fluctuation in 
steam pressure is less than 10 pounds per square inch and the moisture 
in the steam is practically constant, this apparatus is said to register 
within two per cent of the actual weight flowing. 

The Burnham steam meter,. Fig. 422a, a recently patented device, 
offers the advantages of low first cost and simplicity of installation. 
This apparatus is attached to the steam pipe in a manner similar to a 
simple hydrostatic lubricator and occupies about the same space. 
It is based on the principle of the pitot tube, and the weight of steam 
flowing per unit of time is read from a graduated scale. Half-inch 
pipe fittings are used in connecting up. 

In Europe the principles of the Venturi meter have been success- 
fully applied to the measurement of steam. jfGluckauf, Dec. 9, 1905.) 

Steam Meters: Proc. A.S.M.E., Mid.-Nov., 1909, p. 1239. 





DIFFERENTIAL DRAFT GAUGE 



Fig. 423. Different Forms of Manometer Pressure Gauges. 



SIMPLE O TUBE 



406a. Pressure Gauges. — The Bourdon 
type of gauge, either autographic or indi- 
cating (Fig. 424), is the most familiar and 
satisfactory means of measuring pressures 
up to 1500 pounds per square inch or 
more, although diaphragm gauges are also 
used and both are employed as vacuum 
gauges. For the latter purpose, however, 
the mercurial vacuum gauge has the ad- 
vantage of greater accuracy and is not 
subject to derangement. Bourdon gauges 
should be frequently standardized by com- 
parison with a gauge of known accuracy, a mercury column, or a gauge 
tester. 

For measuring very low pressures, such as are found in boiler flues 




Fig. 424. Bourdon Pressure 
Gauge. 



736 STEAM POWER PLANT ENGINEERING 

or gas mains, indicating or recording diaphragm gauges may be had, but 
some form of U tube manometer is generally employed, the design best 
adapted to the purpose depending upon the accuracy required. The 
simple U tube (Fig. 423) when filled with mercury may be used for 
pressures limited only by the inconvenience due to length of tubes, or, 
with water as the fluid, for pressures only a fraction of an ounce per 
square inch. Where greater accuracy is required than can be 
obtained with the simple U tube, some modification may be employed, 
such as the Eames draft gauge with one inclined leg which magnifies 
the reading several times. A form of sensitive gauge is sometimes 
used which depends upon the use of two fluids of different specific 
gravity, as oil and water. 

Pressure Gauges, General References: Mech. Engr., Aug. 17, 1907; Am. Elecn., 
July, 1901; Engng., Aug. 23, 1907; Elec. World, Feb. 2, 1907, p. 258; Power, March, 
1905, p. 184. 

Recording Pressure Gauge: Trans. A.S.M.E., 11-225, 14-325; Elec. World, 
April 28, 1906, p. 886. 

Draft Gauges: Trans. A.S.M.E., 21-123; Engr. U.S., Feb. 15, 1907, p. 218; 
Mech. Engr., Oct. 27, 1906. 

407. Measurement of Temperature. — For power-plant purposes 
mercurial thermometers are most convenient for measuring tempera- 
tures up to 400 degrees F., and are inexpensive. For higher tempera- 
ture, up to say 800 degrees F., they are also adapted, but must be 
made of special glass and the space above the mercury filled with 
nitrogen under pressure to prevent vaporization of the mercury. Such 
thermometers must be used intelligently, and should be standardized 
from time to time, since they are subject to considerable change. 
The Bureau of Standards at Washington, D.C., is prepared to furnish 
certificates for which a nominal charge is made. 

Fig. 425 shows a form of thermometer which is much used where a 
continuous autographic record is required. It depends for its oper- 
ation upon the pressure produced by a fluid, liquid or gaseous, 
contained in a small bulb and exposed to the temperature to be 
measured. The pressure is transmitted to the recording mechanism 
through a flexible capillary tube which may be of . considerable length. 
Such thermometers are suitable for feed water, flue gas, and tempera- 
tures not exceeding 1000 degrees F. 

Fig. 426 illustrates a form of electrical pyrometer employing thermo- 
couples which has come into wide use as a reliable means of measur- 
ing temperatures up to 2600 degrees F. The couples most frequently 
used are composed of platinum and platinum-rhodium, platinum and 
platinum-iridium, copper and copper-constantan, and copper and nickel, 



TESTING AND MEASURING APPARATUS 



737 



the first named being adapted to the higher ranges of temperature. The 
electro-motive force set up, when the thermo-j unction is heated, is pro- 
portional to the temperature and is measured by means of a sensi- 




W//////MM? 




HH/WWM1L 



Fig. 425. Bristol Recording Pyrometer. 




Fig. 426. Bristol Thermo-Electric Pyrometer. 

tive millivoltmeter which is usually graduated to read temperature 
directly. Thermo-couples may be made to give an autographic record 
by means of a thread recorder. 



738 



STEAM POWER PLANT ENGINEERING 



Fig. 427 shows the element of an electrical thermometer based upon 
the change in resistance of a platinum wire when subjected to change 
in temperature. The resistance, in terms of temperature, is measured 
by a Whipple indicator, a convenient and portable form of Wheat- 




Fig. 427. Element for Callendar Resistance Pyrometer. 

stone bridge, or may be autographically recorded by means of a Callen- 
dar recorder. Resistance thermometers of this type are very sensitive 
and accurate, not easily deranged, and are limited in range only by the 
fusing points of the platinum and the porcelain protecting sheath. 

For higher temperatures and for obtaining the temperatures of 
inclosed spaces above about 900 degrees F., such as boiler furnaces, 
annealing ovens, and kilns, various forms of optical and radiation 
pyrometers have been devised. In such devices no part of the instru- 



OIFFUSING GLASS 



FLAME 
GAUGE 



l^AMYL-ACETAT 
LAMP 




Fig. 428. Wanner Optical Pyrometer in Position for Standardizing. 

ment is exposed to the temperature to be measured and hence suffers 
no injury from this cause. Optical pyrometers are based upon the 
measurement of the brightness of the hot body by comparison with 
a standard. The Wanner optical pyrometer is shown in Fig. 428. 



TESTING AND MEASURING APPARATUS 



739 



After standardizing by comparison with an amyl-acetate lamp, it is 
only necessary to focus the instrument upon the source of heat to be 
measured and the temperature is read on the graduated scale. 



TABLE 109- 

TYPES OF THERMOMETERS IN GENERAL USE. 









Range in Degrees F. 


Principle of Operation. 


Type. 


for which they 








can be used. 


Expansion 


.... Those depending on the 


Gas 


— 400 to +2900 

— 35 to +950 




change in volume or 


Mercury, Jena glass, 




length of a body with 


and nitrogen 






temperature. 


Glass and petrol ether. 


— 325 to +100 






Unequal expansion of 


to 950 






metal rods. 




Transpiration and 
cosity. 


vis- Those depending on the 
flow of gases through 


The Uehling 


to 2900 








capillary tubes or small 








apertures. 






Thermo-electric . . 


.... Those depending on the 
electro-motive force 
developed by the dif- 
ference in temperature 
of two similar thermo- 
electric junctions op- 
posed to one another. 


Galvanometric 


— 400 to +2900 


Electric resistance. 


. . . .Those utilizing the in- 


Direct reading on indi- 


— 400 to +2200 




crease in electric resist- 


cator or bridge and 






ance of a wire with 


galvanometer. 






temperature. 






Radiation 


, . . . Those depending on the 


Thermo-couple in focus 
of mirror. 


300 to 4000 










bodies. 


Bolometer 


— 400 to Sun 


Optical 


. . . . Those utilizing the 


Photometric compari- 
son. 






change in the bright- 






ness or in the wave 


Incandescent filament 


1100 to Sun 




length of the light 


in telescope. 




emitted by an incan- 


Nicol with quartz plate 






descent body. 


and analyzer. 




Calori metric 


. . . .Those depending on the 


Platinum ball with 
water vessel. 


32 to 3000 




specific heat of a body- 






raised to a high tem- 








perature. 






Fusion 


. . . . Those depending on the 


Alloys of various fusi- 
bilities. 


32 to 3350 




unequal fusibility of 






various 'metals or 








earthenware blocks of 








varied composition. 







Radiation pyrometers depend upon the measurement of the heat 
radiated from the hot body. The Fery radiation pyrometer, Fig. 429, 



740 



STEAM POWER PLANT ENGINEERING 



is the best-known instrument of this type. When focused upon the 
source of heat a cone of rays of definite angle is reflected by means of 
the mirror upon a thermo-couple located in its focus. The electro- 
motive force set up is measured in terms of the temperature of the 



TO GALVANOMETER 




Fig. 429. Fery Radiation Pyrometer. 



source of heat by a milli volt meter. Neither the couple nor any 
part of the instrument is ever subjected to a temperature much above 
150 degrees F. The indications are practically independent of the 
distance from the source of heat, and the range is without limit. 

Table 109 embodies in outline the principles and temperature ranges 
of the various types of thermometers in use. Temperature ranges 
verified by U. S. Bureau of Standards. 

Indicating and Recording Thermometers, Expansion Type: Sci. Am. Sup., 
Dec. 16, 1905; Trans. A.S.M.E., 22-143; Jour. Am. Chem. Soc, 16-396; Jour. Soc. 
Chem. Ind., 13-61; Philosoph. Mag., 50-251, 1900. 

Indicating and Recording Pyrometers, Thermo-Electric: Jour. West. Soc. Engrs., 
Sept., 1907; Cassier's Mag., Aug., 1905; Elec. Rec., Jan. 12, 1901; Elec. Chem. and 
Met., June, 1901. 

Indicating and Recording Pyrometers, Electric-Resistance: Engng., May, 1899; 
Jour. Chem. Soc, 1890, 1895; Jour. Iron and Steel Inst., 1892; Pro. Royal Inst., 
Vol. XVI, 1901; Bureau of Standards, 3-641, 1907; Electrician, March 17, 1905, 
p. 880. 

Indicating and Recording Pyrometers, Optical: Elecn., Lond., Aug. 17, 1906; 
Am. Mach., Vol. 28, 160-29; Sch. of Mines Quarterly, April, 1907; Bulletin No. 2, 
Bureau of Standards, Wash., 1905; Jour, de Phys., Serjt., 1904; Engng., Sept. 6, 1907, 
Oct. 18, 1907; Cal. Jour, of Tech., Aug., 1907; Bureau of Standards. Bulletin No. 2; 
Iron Age, 73-24. 

Miscellaneous References: Engng. Times, March, 1904; Engng., Feb. 17, 1903, 
March 6, 1904; Sci. Am. Sup., July 22, 1905; Min. Rept., Aug. 8, 1901; Iron Age, 
Feb. 7, 1907; Iron and Coal Tds. Rev., May 10, 1907; Am. Elecn., May, 1904 ; Physical 
Rev., 8-193; Roy. Soc. of Lond., 66-86, 1900. 



TESTING AND MEASURING APPARATUS 



741 



408. Power Measurements. — The indicated horse power of recip- 
rocating engines is usually obtained by means of the steam-engine 
indicator. There are several reliable types to be had, including the 
continuous indicator, which permits of several diagrams being taken 
successively on the same paper. Among other devices may be 
mentioned mean pressure indicators and those giving the horse power 
directly. 

The developed horse power is determined by some form of absorp- 
tion dynamometer. For description of such dynamometers see Appen- 
dix C, article XV, A.S.M.E. code for conducting steam-engine tests. 

Power Measurements: Trans. A.S.M.E., 13-531; Am. Mach., Vol. 30, No. 27, 
Vol. 31, No. 5; Mechanical Engr., Feb. 23, 1907; Engng., June 14, 1907, p. 768. 

Indicators, Continuous: Trans. A.S.M.E., 18-1020; Power, Jan., 1907, p. 26. 

Prony Brakes: Trans. A.S.M.E., 15-62; Am. Mach., July 27, 1905, p. 127; Eng. 
News, Vol. 44, p. 216. 

Water Absorption Dynamometers: Prac. Engr., Sept. 14, 1906; Trans. A.S.M.E., 
11-958; Eng. News, Vol. 51, p. 475; Prac. Engr., Sept. 14, 1906, p. 326. 




Fig. 430. Orsat Apparatus for Flue Gas Analysis. 



409. Flue-Gas Analysis. — The simplest device for the analysis of 
flue gases is the Orsat apparatus (Fig. 430). In this apparatus a 



742" 



STEAM POWER PLANT ENGINEERING 



measured volume, representing an average sample of the gas, is forced 
successively through pipettes containing solutions of caustic potash, 
pyrogallic acid, and cuprous chloride in hydrochloric acid, respectively, 
thus removing the carbon dioxide, the oxygen, and the carbon monox- 
ide, the contraction of volume being measured in each case. 

Orsat Apparatus: Trans. A.S.M.E., 18-901 ; Steam Boilers, Peabody and Miller, 
Chap. II; Power, Aug., 1907, p. 532; Engr. U.S., Jan. 1, 1907, p. 71. 

For most practical purposes it is sufficient to determine the carbon 
dioxide. A simple and efficient device for continuously indicating 
the per cent oT C0 2 is Arndt's econometer, Fig. 431. This apparatus 



rM±M 




Fig. 431. Arndt's Econometer. 



is a gas-weighing balance consisting essentially of a sensitive beam 
from one end of which is suspended a glass globe, closed at the top and 
open at the bottom, of about one pint capacity, and from the other 
end a compensating rod and scale pan. When not in operation the 
globe is filled with air and the scale pan and globe are in perfect 
balance, the indicator pointing to zero. When in operation the flue 
gases, thoroughly dried and filtered, are introduced in a continuous 
flow into the body of the hollow globe by means of a glass tube. 
The larger the per cent of C0 2 present in the contents of the globe 
the greater will be the deflection of the pointer, since C0 2 is about 



TESTING AND MEASURING APPARATUS 



743 



50 per cent heavier than atmospheric air. The scale is graduated 
to read from to 21 per cent C0 2 , and the results obtained check 
closely with those of the Orsat apparatus. Arndt's econometer is 
not portable, though it may be placed almost anywhere where it can 
be easily seen by the fireman. When there are a number of boilers, 
and it is not desired to have a separate instrument for each, the eco- 
nometer is connected with the breeching of each boiler by suitable piping, 
the gas from one boiler at a time being analyzed. 

For descriptive details see circular issued by Joseph Wickes, 106 
Fulton Street, New York. 



MM///////7777n///// ///S//////A 




<zs 



3 C*S INLCT 



The Ados recorder, Fig. 432, Sarco recorder, Fig. 433, and Uehling 
composimeter are well-known types of instruments which give con- 
tinuous autographic records of the percentage of C0 2 on clock-driven 
charts. These devices are very satisfactory, but are rather expensive 
and usually beyond the appropriation of small boiler plants. 

The Sarco C0 2 recorder operates as follows: Gas is drawn con- 
tinuously from the flue through a f-inch pipe by means of the aspi- 



744 



STEAM POWER PLANT ENGINEERING 



rator Q, and enters the apparatus through the tube D, flowing as indi- 
cated by the arrows through C and E. The aspirator requires from 6 
to 8 gallons of water per hour, which is discharged into the pressure- 
regulating tank L. The greater 
portion overflows through R, 
and the balance is caused to 
issue in a fine stream by adjust- 
ment of cock S into H and G 
and thence to an upper com- 
partment of vessel K, which it 
gradually fills, compressing the 
air which it contains and trans- 
mitting this pressure to the 
lower chamber through a con- 
necting tube. The lower cham- 
ber is filled with a solution of 
1 part glycerine to 3 parts 
water, which is driven out into 
the calibrated vessel C. When 
it has risen to the height of the 
inlet and outlet the flow of gas 
is interrupted and a portion is 
trapped. By the time the lower 
end of the center tube in C is 
sealed off the excess of gas has 
been forced out against the 
slight resistance of the elastic 
bag P, and the vessel contains 
exactly 100 c.c. at atmospheric 
pressure. During this time the 
aspirator is drawing the flue gas 
through the seal F. As the 
liquid rises higher in C, the en- 
trapped gas is forced through 
the small tube Z into A, which 
is filled with a solution of caus- 
tic potash. Here the C0 2 is absorbed and the potash solution 
is forced into vessel B, which has an outer jacket filled with 
glycerine supporting a float N suspended from the pen- gear M. 
A thin tube through the float keeps the air in B at atmos- 
pheric pressure. As the float rises the pen-lever swings upward, 
carrying the pen Y, which makes a vertical line upon the clock-driven 




Fig. 433. Sarco CO, Recorder. 



TESTING AND MEASURING APPARATUS 



745 



recording drum. The quantity of potash solution displaced, and con- 
sequently the lift of the float, is dependent upon the amount of C0 2 
absorbed from the 100 c.c. of flue gas. The chart is graduated to 
indicate directly the percentage absorbed. By the time this opera- 
tion has been completed water has filled tank K and risen into the 
siphon G, which, upon starting, very rapidly empties the tank and 
allows the liquid from C to return to the lower compartment. The 
float returns to its original position and the remaining gas passes out 
through E. 

Ados C0 2 Recorder: Engng., Jan. 12, 1906; Sci. Am. Sup., Dec. 22, 1906. 
Sarco C0 2 Recorder: U.S. Engr., Nov. 1, 1907, p. 1001. 
Uehling Composimeter: Power, June, 1907, p. 404. 
American C0 2 Indicator: Power, Dec, 1907. 

Flue Gas Analysis, Miscellaneous Apparatus: Power, April, 1907, p. 243; Engr. 
U.S., Jan. 1, 1907, p. 71; Elecn., Lond., Nov. 16, 1906. 

410. Moisture in Steam. — Several forms of calorimeters are avail- 
able for determining the quality of steam. The simplest as well as 
the most satisfactory, if the percentage of entrained moisture is not 
beyond its range, is the throttling 
calorimeter, Fig. 435. In this device 
the sample of steam, .which is taken 
from the steam pipe by means of the 
perforated nipple, is allowed to expand 
through a very small orifice into a 
chamber open to the atmosphere. The 
excess of heat liberated serves first to 
evaporate any moisture present and 
then to superheat the steam at the 
lower pressure. From the observed 
temperature and pressures it is easy 
to calculate, with the aid of steam 
tables, the percentage of moisture in 
the original sample. 

The limit of the throttle calorimeter 
depends upon the steam pressure and 
is about 3 per cent of moisture at 80 
pounds pressure and about 5 per cent 
at 200 pounds. For steam containing 
greater percentages of moisture the 
separating calorimeter, Fig. 434, is 

sometimes used. This instrument is virtually a steam separator and 
mechanically separates the moisture from the sample of steam. 



SEPARATOR 




GRADUATED 
SCALE 



STEAM JACKET 



DISCHARGE 
ORIFICE 



Fig. 434. Carpenter Separating 
Calorimeter. 



746 



STEAM POWER PLANT ENGINEERING 



- THERMOMETER 




TO ATMOSPHERE -" 

Fig. 435. A Typical Throttling Calorimeter. 



THERMOMETER WELL 



THROTTLE W«LVt 



(ERMOMETEfl WtLC 



rMOSPHERIC D(SCH»RC« 




r\ 



w 



Fig. 435a. Ellison Universal Steam Calorimeter. 



TESTING AND MEASURING APPARATUS 



747 



The water thus separated collects in a reservoir provided with gauge 
glass and graduated scale, while the dry steam, passes through an 
orifice to the atmosphere. The weight of dry steam per unit of time 
is indicated on the gauge, calculated according to Napier's rule, or may 
be determined by condensing and weighing. The accuracy of the 
moisture determination is greatly affected by the difficulty of obtain- 
ing true samples of steam containing large percentages of moisture. 

Fig. 435a shows the Ellison universal steam calorimeter, which 
combines the superheating and throttling principles and is adapted to 
steam of any degree of wetness. The separating chamber is provided 
with a gauge glass, not shown, for indicating the weight of water which 
accumulates only when the steam is too wet to be superheated. 

Throttling Calorimeters: Power, Dec, 1907, p. 891; Trans. A.S.M.E., ,17-151, 
175, 16^48; Engr. U.S., Feb. 15, 1907, p. 219. 

Separating Calorimeters: Trans. A.S.M.E., 17-608; Engr. U.S., Feb. 15, 1907, 
p. 219. 

Universal Calorimeter: Trans. A.S.M.E., 11-790. 

Thomas Electrical Calorimeter: Power, Nov., 1907, p. 791. 

411. Fuel Calorimeters. — The analysis and heat evaluation of 
fuel require considerable time and skill and much costly apparatus, 



INSULATION 

BOMB 

PLATINUM PAN 

WATER 

ELECTRODE 

IGNITION WIRE 

STIRRING DEVICE 

SUPPORT FOR STIRRER 

SENSITIVE THERMOMETER 

OXYGEN TANK 




Fig. 436. Mahler Bomb Calorimeter. 



hence in most power plants it is customary to depend upon a specialist 
to whom samples are submitted from time to time. In many large 



748 



STEAM POWER PLANT ENGINEERING 



stations, however, the conditions often warrant the establishment of a 
testing laboratory equipped for the proximate analysis of coal and the 
determination of the calorific value of the solid, liquid, or gaseous fuel 
used. The Mahler bomb calorimeter illustrated in Fig. 436 is the 
most accurate and satisfactory device for solid and liquid fuels but is 
comparatively expensive. The instrument consists of a steel shell or 
" bomb " of great strength, lined with porcelain or platinum, into 
which a weighed sample of the fuel is introduced and burned on a 
platinum pan in the presence of oxygen under a pressure of about 300 
pounds per square inch. The charge is ignited by an electric current. 
During combustion the bomb is submerged in a known weight of 
water which is kept constantly agitated. The calorific value is calcu- 
lated from the observed rise in temperature due to the heat evolved, 
proper corrections being made for the water equivalent of bomb and 
appurtenances, heat given up by the igniting current, and for radiation 
or absorption of heat from the surrounding air. 




COMPRESSED 
FIBER 



STIRRER 



Fig. 437. Parr Fuel Calorimeter. 



The Parr calorimeter, Fig. 437, is an inexpensive instrument, very 
simple in operation, and gives results which are sufficiently accurate 
for all practical purposes. The weighed sample of coal, together with 
a quantity of sodium peroxide which supplies the oxygen for com- 
bustion, is introduced into the cartridge. Means are provided for 



TESTING AND MEASURING APPARATUS 749 

rotating the cartridge when submerged in the calorimeter, the 
attached vanes agitating the water to maintain uniform temperature. 
The charge is fired either electrically or by introducing a short piece of 
hot wire through the conical valve. The calorific value is calculated 
from the observed rise in temperature and the constants of the instru- 
ment. Among other forms of instruments, in more or less general use 
and which give very satisfactory results, may be mentioned the Car- 
penter calorimeter and the Thompson calorimeter. 

Mahler Bomb Calorimeter: Engr. U.S., Jan. 1, 1907, p. 68. 

Parr Fuel Calorimeter: Power, July, 1907, p. 499; Engr. U.S., April 1, 1903; 
Jour. Am. Chem. Soc, 22-246; The Calorific Value of Fuels, Poole; Gas and Fuel 
Analysis, Gill; Eng. Chem., Stillman; Chem. Technology, Groves and Trop. 

Carpenter Coal Calorimeter: Trans. A.S.M.E., 16-1040. 

Thompson Coal Calorimeter: Jour. Soc. Chem. Ind. (1906), 25-409. 

Junkers Gas Calorimeter: Jour. Soc. Chem. Ind. (1895), 14-631 ; Stevens Indicator, 
Jan., 1905, p. 31. 

Comparison of Different Types of Calorimeters: Jour. Soc. Chem. Ind. (1903), 
22-1230. 

411a. Hamler-Eddy Smoke Recorder. — This apparatus consists 
essentially of a small motor-driven vacuum pump, which draws a con- 
tinuous sample of the products of combustion from the uptake breeching 
or stack and discharges it against a paper-covered drum revolved by 
clockwork. The density of the smoke, the time at which visible smoke 
is being emitted and the duration of the smoke production period are 
automatically recorded on the paper by the smoke itself. Before 
reaching the pumps the gases pass through a glass " emergency " con- 
denser and a large portion of the vapor content is removed. The pump 
discharges the partially dried gases against a surface of sulphuric acid 
(which removes the last trace of moisture) and forces the smoke in the 
form of a small jet of dry powder onto the surface of the recording paper. 
The sampling tube leading from the flue to the pump is connected with 
a steam line and is " blown out " each time a card is changed. The 
instrument is very compact and portable and may be placed anywhere 
with respect to the chimney. A number of these appliances in Chicago 
power plants are giving excellent satisfaction. 



CHAPTER XIX. 



TYPICAL SPECIFICATIONS. 



412. Sample Specifications for a Cross Compound Non-Condensing 
Engine.— For and in consideration of the amount and terms named in the letter 
accompanying this specification, and of the same date, we propose to furnish 
f.o.b. cars at our factory, Elizabethport, N.J., for account of The Armour 
Institute op Technology, Chicago, 111., One Ball & Wood Horizontal 
Four- Valve (Corliss) Center-Crank Engine, designed for direct connection 
to a direct-current generator, as follows : 

General Horse power 350 to 375 

Dimensions. Diameter of cylinders H.P., 17; L.P., 27 inches 

Length of stroke 18 inches 

Revolutions per minute 175 to 200 

Governor wheel diameter, 90; face, 21 inches 

Width of belt (if belted) 20 inches 

Diameter of steam pipe 7 inches 

Diameter of exhaust pipe 10 inches 

Crosshead pins 6 in. long, 8 in. diameter 

Crank pins 9 in. long, 9$ in. diameter 

Main bearings 20 in. long, 9 in. diameter 

Wearing surface of crossheads 242 square inches 

Width of engine over all 14 feet 6 inches 

Length of engine over all 20 feet 3 inches 

Weight complete 60,000 pounds 

Rating. The rated power of the engine specified is based on an initial 

pressure of 120 pounds (measured in the cylinder), cutting off at 
about one third stroke in both cylinders, without vacuum, when 
operating at 200 revolutions per minute. 
Fittings. With each engine is furnished the following complete list of 

fittings : 

One extended shaft (omitted if belted engine), 

One sub-base with extension for dynamo (omitted if belted 

engine), 
One self-oiling outboard bearing (omitted if belted engine), 
One throttle valve, 
One lubricating system consisting of pipes, sight feeds, and 

oil reservoir, 
One cylinder lubricator, nickel plated, 
750 



TYPICAL SPECIFICATIONS 751 

Fittings — Continued. 

One set special steel wrenches, 

One socket wrench for piston, 

One socket wrench for connecting rod bolts, 

One steel wrench for hexagon nuts, 

One connecting rod set screw wrench, 

One monkey wrench, 

One spanner for valve stem gland, 

One eye bolt for pillow block cap, 

Two push-off bolts, 

One set grease cups, 

Two oil cups, 

One hand oil pump, 

One set brass oil cups, 

Three nipples for drip pipes, for frame, 

Four nipples for cylinder drips, 3 inches long, 

One nipple for throttle bleeder, 

One globe valve for throttle bleeder, 

Four globe valves for cylinder drips, 

One set foundation bolts, nuts, and plates, 

One template for locating bolts, 

One governor wheel and keys, 

One balance wheel and keys (omitted if direct connected 
engine), 

Packing for piston and valve rods, 

One one-gallon can cylinder oil, 

One one-gallon can engine oil, 

Two cans grease, 

Two wedges for wheels, 
Cylinders. Cylinders are made of hard and close-grained iron, and under 

the influence of oil and wear the walls will rapidly acquire a fine, 
smooth glaze. Radiation is prevented by a thick jacket of 
asbestos cement, outside of which is neatly fitted an orna- 
mented jacket. Openings and globe valves are provided for 
drainage. 
Connecting The connecting rods are of forged machinery steel of low 
Rods. carbon and fitted with heavy straps with keys and bolts for 

adjustment. 

Crank Pin The crank pin boxes are of cast iron lined with Babbitt 

Boxes. metal. 

Crosshead The crosshead boxes are of the best quality of phosphor 

Boxes. bronze. 

Crossheads. The crossheads are of cast iron faced with Babbitt metal on the 
wearing surfaces. The crosshead pins are pressed into the cross- 
heads. 



752 



STEAM POWER PLANT ENGINEERING 



Piston. The piston rods are of special hammered steel, threaded and 

screwed into the crossheads, and locked fast with special nut 
counterbored at the end to cover threads, finished and case- 
hardened. The other ends of the rods will be fitted to the pistons 
with thread and locked with nut. The pistons will be fitted with 
two rings turned eccentric and cut open at the thinnest part, the 
ends being halved so as to lap when in position. 

Valves. Both the admission and exhaust valves are of the Corliss 

pattern. The former are provided with double ports, and are 
actuated from a wrist plate receiving its motion from the governor 
placed in the fly wheel of the engine. This governor controls the 
valves of both the high and low-pressure cylinders and possesses 
a range of cut-off from to about f stroke. 

The exhaust valves are driven from a wrist plate through an 
adjustable eccentric by which any desired degree of compression 
can be obtained. 

Speed. The use of Corliss valves, arranged as described in the foregoing 

paragraph, permits an increased speed over the common type 
of Corliss engine with releasing gear, and while yielding the 
same economy dispenses with many working parts, and, what 
is more important, with the large and cumbersome fly 
wheel which has so often proved a source of danger in slow- 
speed engines. 

The frame is proportioned for great strength and the metal is 
placed where it is most needed. An oil groove is cast around the 
bottom of the frame to protect the foundation. 

Main bearings are fitted with removable Babbitt shells which 
can be replaced when necessary. Special care is taken to have 
these bearings of ample length to support the wheels and stand 
the strain of power transmission or the weight of armature when 
direct connected. In the latter case a self-oiling outboard bearing 
is provided to carry the outer end of shaft. 

Guides, The guides are known as the locomotive pattern and are inter- 

changeable. They are carefully scraped to surface plates and 
provision made for taking up wear. 

Crank Shaft. The crank shaft is of the best quality of steel, being carefully 
counterbalanced by cast-iron disks in which the necessary weight 
is placed. In direct connected engines this shaft is either 
extended in one piece to carry the armature or made in two 
pieces and coupled. 

Governor. The governor is of the inertia type and has a swinging eccen- 

tric, the eccentric center moving across the end of the shaft about 
an outside point, and giving a lead which varies with the point of 



Frame. 



Main 
Bearings. 



TYPICAL SPECIFICATIONS 



753 



Governor — Continued. 

cut-off from a maximum, at the latest point, to zero, when the 
governor weights occupy their extreme outward position. Alter- 
ation in speed is obtained by changing the amount of weight in 
the pockets of the lever arm. 

Balance. The balance wheel (in the case of belted engine) is made with 

a flanged rim and with a split hub, the hub being secured to the 
shaft by a bolt. The other keys are square, with parallel sides, 
and are inserted without driving. 

Oiling. The oiling system consists of a simple oil reservoir which sup- 

plies oil through a system of pipes to the points of the engine 
needing lubrication. After fulfilling its functions this oil is drained 
and can be used anew. This does away with the old cumbersome 
oil-cup system and has the great advantage of delivering clean oil 
to the engine. 

Guarantees. Material. We guarantee that the material and workmanship 
are of the best and that all working parts having flat surfaces are 
scraped to surface plates. 

Regulation. That the engine shall regulate within 2 per cent 
under changes of load within the range of the governor, and that 
no reduction of boiler pressure shall reduce the speed until the 
latest point of out-off is reached. 

Steam Consumption. That the steam consumption, when the 
engine is developing its rated power at 125 pounds pressure 
and no vacuum, shall not exceed 22 pounds of dry steam per 
indicated horse power per hour; that the clearance shall not 
exceed 8 per cent. 

Drawings. With the engine is furnished a drawing showing its details, 

together with foundation plans. 

Preparation Every engine is completely erected at our works before ship- 

for Ship- ment. The castings are rubbed smooth, carefully filled, and the 

ment. engine given two good coats of standard shop color. All bright 

parts are carefully protected against corrosion. The engine is 

dismantled, the small parts being boxed, and in the case of export 

shipment the larger pieces crated. 

Erection. Full drawings and directions for erecting the engine will be 

furnished. Template, foundation bolts, nuts, and plates to be 
shipped in advance if necessary, and by freight unless otherwise 
directed. 

If requested we will furnish the services of an expert to superin- 
tend the erection of this engine at the rate of $5 per day added 
to his traveling expenses and board, the purchaser to furnish all 
laboring help. 



754 



STEAM POWER PLANT ENGINEERING 



Terms. One-half cash on presentation of bill of lading, balance on 

completion of erection. 

The title to the apparatus herein sold shall not pass from The 
Ball and Wood Company until all payments hereunder (including 
deferred payments, if any) shall have been fully made in cash. 
The purchaser agrees to do all acts necessary to perfect and main- 
tain such retention of title in the said Company. All previous 
communications between the parties hereto, verbal or written, are 
hereby abrogated and withdrawn, and this proposal, when duly 
signed and approved, constitutes the agreement between the 
parties hereto, and no modification of this accepted agreement 
shall be binding upon the parties hereto or either of them unless 
such modification shall be in writing, duly accepted by the pur- 
chaser and approved by an executive of the Company. 

Limit. Prices subject to revision after thirty days. Delivery subject 

to strikes, accidents, or causes beyond our control. 

413. Specifications for Horizontal Tubular Steam Boiler.* — The 

following specifications for one 54-inch horizontal return tubular steam 
boiler, pressure 125 pounds, were prepared by the Hartford Steam 
Boiler Inspection and Insurance Company for the Armour Insti- 
tute of Technology, Chicago. 

The boiler to conform to the following conditions and requirements : 

Type and It is to be of the horizontal tubular type, set with overhanging 

General front, and all parts and pieces are to be designed accordingly. 

Dimensions. ft j s to be 17 feet 2 inches long, outside, and 54 inches in dia- 
meter, measured on the outside of the smallest ring of plates. 
Heads are to be 16 feet inches apart, outside. 
Materials: Shell plates are to be three-eighths of an inch thick on the 

Quality, edges, of open-hearth fire-box steel, having a tensile strength of 

Thickness, not less than 55,000 pounds nor more than 62,000 pounds per 
and Tests. square inch of section, and an elastic limit of not less than half 
the tensile strength, with not less than 56 per cent of ductility, as 
indicated by contraction of area at point of fracture under test, 
and by an elongation of 25 per cent in a length of 8 inches. 

Heads are to be one-half of an inch thick, of best open-hearth 
flange steel. All plates, both of shell and heads, are to be plainly 
stamped with name of maker, brand, and tensile strength ; brands 
so located that they may be seen on each plate after the boiler is 
finished. 

Each shell plate is to bear a coupon which shall be sheared off, 
finished up, and tested by, or for, the maker of the boiler, at his 
expense. Each coupon is to fulfill the foregoing requirements as 

* Drawings have been omitted. 



TYPICAL SPECIFICATIONS 



755 



Materials: Quality, Thickness, and Tests— Continued. 

to strength and ductility, and stand bending down double when 
cold, when red hot, and after being heated and quenched in cold 
water, without signs of fracture. There is not to be more than 
0.035 per cent of sulphur, nor more than 0.035 per cent of phos- 
phorus in the chemical composition of the plates and heads. All 
plates failing to pass these tests will be rejected. All tests and 
inspections of material may be made at the place of manufacture 
prior to shipment. Certified copies of report of tests must be 
sent to the Hartford Steam Boiler Inspection and Insurance 
Company, Hartford, Conn. 

Riveting. The longitudinal seams are to be of the double-riveted butt- 

joint type with double covering strips. They are to be arranged 
to come well above the fire line of the boiler, and break joints 
in the 3 ring courses in the usual manner. The plates are to be 
planed on the caulking edges before rolling. 

All dimensions and proportions are to be shown on accompany- 
ing drawing No. 1502. 

The girth seams are to be of the single-riveted lap-joint type; 
rivets to be of same size as those in longitudinal seams, and 
pitched 2£ inches apart from center to center; the distance from 
center of rivet to the edge of the plate to be equal to 1^ times 
the diameter of rivet hole. 

The rivet holes are to be either drilled in place, or punched at 
least one-quarter of an inch less than full size ; if the latter method 
is used, the plates, after punching, are to be rolled and bolted 
together, and the rivet holes drilled in place one-sixteenth of an 
inch larger than the diameter of the rivets. The plates are then 
to be disconnected. All burrs are to be removed from the edges 
of the holes. Should any holes be in the least out of true, they 
are to be brought in line with a reamer or drill; if a drift-pin is 
used for this purpose the boiler will be rejected. 

All rivets are to be driven by hydraulic pressure, wherever 
possible, and allowed to cool and shrink under pressure. This pres- 
sure is to completely fill the rivet holes, producing a tight joint. 
Rivet Ham- The rivets are to be of the best quality of iron or soft steel, 
mer Tests, capable of being hammered flat, when cold, to a thickness of 
one-half their original diameter, or when hot, to one-third their 
original diameter, without showing signs of fracture. In the 
absence of physical test, it is understood that the contractor 
guarantees the above quality of rivets. 
Braces, There are to be 20 braces 1^ inches in diameter in the boiler, 

10 above the tubes on front head, and 10 on rear head, of the 
crow-foot form, arranged as shown on drawing. None of them 
is to be less than 3 feet 6 inches long, and each is to be fastened 



756 



STEAM POWER PLANT ENGINEERING 



Braces — Continued. 

to shell and heads by two seven-eighths inch rivets at each end; 
or solid steel, diagonal braces of approved pattern, and of equal 
strength to the former, may be used. Care is to be exercised in set- 
ting them that they may bear uniform tension. Crow-foot braces 
may be flat in body, if of equal strength to those specified above. 
Braces There are to be 4 braces below the tubes in the boiler. Two 

below of these are to be through braces extending from head to head. 

Tubes. Each brace is to be 1^ inches in diameter, with a fork formed on 

rear and secured with a 1^-inch turned bolt and nut to a crow- 
foot securely riveted to rear head; these are the inner or central 
braces. The front end of brace is to be upset to a diameter of 1 1 
inches, threaded and secured to front head with a nut and washer 
on both the inside and outside of head. 

The 2 remaining braces are each to be 1^ inches in diameter, 
and secured to rear head in same manner as the through braces; 
the front end of the brace is to be extended forward, fitted to side 
of shell, and riveted there with two 1-inch rivets. All to be 
substantially as shown on accompanying diagram of tube head 
No. 2431. 
Tubes : Size, There are to be 36 lap-welded or seamless-drawn tubes, of the 
Number, and best quality with regard to tensile strength and ductility. They 
, Arrange- are to be round, straight, and free from all surface defects, prop- 
ment. erly annealed on their ends, and guaranteed by the manufac- 

turers to have been tested to at least five hundred (500) pounds 
per square inch internal hydrostatic pressure. Each tube is to 
be 4 inches in diameter, 16 feet inches long, and not less than 
standard thickness, set in vertical rows, with a clear space 
between them, vertically and horizontally, of 1 inch, except the 
central vertical space, which is to be 2 inches, as shown on accom- 
panying diagram of tube head No. 2431. 

Holes for tubes are to be neatly chamfered off on outside. 
Tubes to be set with a Dudgeon expander, and beaded down at 
each end. Tube holes may be drilled and reamed, or may be 
punched one-quarter inch less than full size, then rose bitted to 
exact diameter. 
Manholes. There are to be two manholes, one 11 x 15 inches, with pressed 

steel frame, double riveted to inside of shell on top, and one 
10 x 15 inches, flanged in front head below tubes, with suitable 
plates, yokes, and bolts, the proportions of the whole such as will 
make them as strong as any portion of the shell of like area. 

Boiler The boiler to be suspended from steel I beams, 6 inches deep, 

Supports. 12i- pounds per foot, by means of eye or U bolts and plate loops. 

There are to be 6 loops, 2 on each side of the boiler, securely 

riveted to boiler shell. The I beams are to be supported on cast- 



TYPICAL SPECIFICATIONS 



757 



Boiler Supports — Continued. 

iron columns of square or rectangular section 6 inches square, 
three-quarters inch thick. Each pair of beams is to be connected 
together, 3 inches apart, by tie-bolts and cast-iron separators ; one 
separator near each end, and others at intervals of about five 
feet. The top and bottom flanges of columns are to be faced true. 
The whole system of suspension is to be made in the best man- 
ner, properly arranged to allow free expansion of the boiler, 
securely held and supported in every direction, amply strong in 
every part, and finished complete. 

Nozzles. There are to be two heavy cast nozzles, made of gun-iron or 

steel, one 4 inches internal diameter for steam-pipe connection, 
and one 6 inches internal diameter for safety-valve connection, 
each accurately squared on top flange, and securely riveted to 
boiler on top. Forged or pressed steel pipe flanges may be used 
in place of nozzles. 

The flanges of the nozzles to correspond in diameter and thick- 
ness with standard extra heavy pipe fittings. 

Smoke There is to be an opening 10 by 62 inches cut out of front 

Opening. connection on top for attachment of uptake or flue. 

Feed Pipe. There is to be a hole tapped in front head for a brass bushing, 

3 inches above the top of upper row of tubes, and 16 inches from 
center of boiler, on left-hand side, for 1^-inch feed-pipe connection. 
The bushing is to be not less than 2 inches long, to permit both 
the external and internal feed pipes to be screwed into it not 
less than seven-eighths inch. 

Also furnish and put in a l|-inch feed pipe extending from 
front head back to within two feet of rear head of boiler, thence 
across the boiler to near shell on right-hand side. On this end 
place an elbow with the outlet pointed down as shown on draw- 
ings. Feed pipe is to be properly hung from the braces. 

Blow-off There is to be an extra heavy pressed steel pipe flange, riveted 

Pipe Con- to bottom of shell, near rear end, and tapped to receive a 4-inch 
nection. extra heavy blow-off pipe. Blow-off valve and fittings to be 

extra heavy. 
Fusible There is to be a fusible plug in rear head, two inches above 

Plug. top of upper row of tubes. 

Fittings. There is to be furnished one pop safety valve 3 inches in 

diameter, one 6-inch steam gauge, three three-quarters-inch gauge 
cocks, and one three-quarters-inch gauge glass 12 inches long, all 
to be of approved pattern, and the necessary holes to be made for 
their proper connection. If combination water column is used, 
the steam and water connections between it and the boiler must 
be made by pipe not less that 1^ inches in diameter. 



758 



STEAM POWER PLANT ENGINEERING 



Castings There is to be furnished a substantial cast-iron front, with all 

for Setting, necessary anchor bolts, 10 feet long, closely fitting front connec- 
tion doors with suitable fastening to prevent warping, closely 
fitting furnace doors with liner plates, rear connection door 
16 x 24 inches, with liner plates, grate bars for grate, pattern to 
be selected by purchaser of boiler, 54 inches long by 48 inches wide, 
with suitable bearer bars for same, arch bars for rear connection, 
and all buckstaves, with the necessary bolts or tie rods, and all 
other castings or ironwork of any description necessary for the 
proper construction and setting of the boiler complete. 
In General. The intent of the foregoing specification is to provide for 
material and workmanship of the best quality, and any details of 
equipment not mentioned in this specification, or not shown on the 
drawings, but necessary for the proper completion of the boiler 
ready for operation, and to be hereafter contracted for, must be 
of equally good quality. 

The size and description of parts are to conform substantially 
to the details of the accompanying plan, and the boiler, complete, 
is to be delivered at and all of the material and workman- 
ship is to be subjected to the inspection and approval of the 
Hartford Steam Boiler Inspection and Insurance Company. 

414. Specifications for Barometric Condenser and Auxiliaries. — 

The following sample specifications for a barometric condenser will give 
some idea of the various items called for in the purchase of a condenser 
and appurtenances, the italicized items being specified by the purchaser. 



" We submit herewith our tender for one condensing 

plant as follows: 

Rated The condenser and auxiliary machinery will have sufficient 

Capacity. capacity to condense 250 pounds of steam per minute (equivalent 

to the steam exhausted from engines developing 1000 horse 

power on a basis of 15 pounds of steam per horse power per hour) 

when supplied with cooling water at a temperature of 70 degrees 

Fahrenheit, and maintaining a vacuum at the condenser of 26 

inches of mercury. 

Capacity un- The plant will also have to condense the quantities of steam 

der Variable unc [ er the varying conditions as stated below : 

Conditions. 



Steam Condensed per 
Minute, Pounds. 


Temperature of Cooling 
Water, Degrees F. 


Vacuum Maintained at 

Condenser, Inches of 

Mercury. 


250 
300 
340 
250 
290 


70 
70 
70 
80 
80 


26 
25 
24 
25 
24 



TYPICAL SPECIFICATIONS 



759 



Quantity of 

Cooling 

Water. 

Apparatus 
Furnished. 



Price and 
Delivery. 

Terms. 



Superin- 
tendence. 

Steam 
Pressure. 



Head 

Pumped 

against. 

Power 
Consump- 
tion of 
Auxiliaries. 



The volume of cooling water required when the condenser is 
working under the above conditions will be from 550 gallons to 
650 gallons per minute. 

The apparatus to be furnished by us will consist of : 

One cast-iron condensing vessel, complete with barometric 
tubes and foot valves. 

One automatic vacuum regulator. 

A structural steel framework for supporting the condensing 
vessel. 

One positive rotary pump for supplying the cooling water to 
the condenser. 

One "dry" air pump. 

One horizontal steam engine, arranged to drive the water pump 
by belt and the air pump direct, the latter placed tandem to the 
engine ; the engine is to be fitted with a suitable governor arranged 
for variable speeds. Purchaser to furnish belt. 

Two pulleys, one for the engine and one for the water pump. 

One vacuum gauge. 

Four thermometers. 

We do not include any steam, air, or water pipes, valves nor 
foundation bolts, but will furnish plans showing suitable founda- 
tions and general arrangement of the machinery. 

Our price for one barometric condensing plant, as described, 
including all royalties, and delivered f.o.b. cars our works, is 



Payments as follows: Monthly payments as the work pro- 
gresses in our shops, less 10 per cent. The retained percentage 
to be paid when the condenser is started in service, provided 
this is done within a reasonable time after completion. 

If desired, we will furnish a competent machinist to superintend 
the erection and starting of the plant, charging extra for his serv- 
ices, 50 cents per hour and his traveling and boarding expenses. 

The engine driving the air and water pumps will be capable of 
starting and operating the plant with 1 00 pounds minimum steam 
pressure, and will be built strong enough to work under 135 
pounds maximum steam pressure. 

The water pump and engine driving same will be designed to 
raise the injection water from the cold well to the condenser, the 
level of the water in the cold well to be not more than 10 feet below 
the level of the water in the hot well. 

The engine driving the air and water pumps will require approxi- 
mately 2 per cent of the main engine steam when operating under 
rated load and with conditions as above stated. 

The rotary pump has a positive displacement (Bibus* patent) 



760 STEAM POWER PLANT ENGINEERING 

Power Consumption of Auxiliaries — Continued. 

and is of substantial construction. All the power for propelling 
the water is transmitted through the main shaft, the office of the 
gears being simply to keep the sealing runner in time with the 
propelling runner. Stuffing boxes are placed between the pump 
chamber and bearings to prevent any grit in the water coming 
into contact with the journals. 

The engine and air pump is of the crank and fly wheel type, 
the engine being fitted with a suitable governor arranged for vari- 
able speeds, and the air pump with the patented slide valve. 

415. Specification for Steam, Exhaust, Water, and Condenser Piping 
for an Electric Power Station.* — It is intended that this specification 
shall cover the complete installation of steam piping, exhaust piping, 
injection and discharge piping, drain, drip, blow-off, and boiler-feed 
piping, water piping, valves, separators, anchors, fittings, etc., sub- 
stantially as shown on the accompanying drawing or hereinafter 
described. 

All the materials used throughout must be the best of their respec- 
tive kinds, subject to the inspection and approval of the engineer of 
the purchaser. 

The entire work provided for in this specification is to be constructed 
and finished in every part in a good, substantial, and workmanlike 
manner, according to the accompanying drawings and this specifica- 
tion to the full intent and meaning of the same, and everything necessary 
for the proper and complete execution of the plans and drawings, 
whether the same may have been herein particularly specified or not, 
or indicated in the plans referred to, to be done and furnished in a 
manner corresponding with the rest of the work as well, as truly, and 
as faithfully as if the same were herein particularly described and 
specifically provided for. 

The engineer shall have full power at any time during the progress 
of the work to reject any materials he may deem unsuitable for the pur- 
pose for which they are intended, or which are not in strict conformity 
with the spirit of this specification. He shall also have the power to 
cause any inferior or unsafe work to be taken down and altered at the 
cost of the contractor. 

It is to be understood that the final inspection and acceptance of 
the work are to take place at the building after erection, and that any 
inspection and acceptance of material and workmanship at the mills,, 
shops, etc., to facilitate the progress of the work, shall not preclude 
rejection at the building if the same be unsuitable. 

* Stevens Indicator, Vol. 18, p. 373, 1901. 



TYPICAL SPECIFICATIONS 761 

Any disagreement or difference between the purchaser and the con- 
tractor, upon any matter or thing arising from this specification or the 
drawings which form a part thereof, shall be referred to the engineer, 
whose decision and interpretation of the same shall be considered final, 
conclusive, and binding upon both parties. 

Risk and Blame. — The contractor is to assume all risks and bear 
any loss occasioned by neglect or accident during the progress of the 
work until the same shall have been completed and accepted by the 
engineer. 

Permits. — The contractor must pay for all permits and inspec- 
tors' fees or any other charges from borough, city, county, or state 
officers. 

Dimensions. — During the progress of the 1 work the contractor will 
be required to keep at the building site a complete set of drawings 
and a copy of this specification. The contractor will consider dimensions 
shown on drawings to be approximate as to location of machinery, 
but sufficiently accurate for the purpose. Absolute dimensions must 
be gotten from the location of the boilers, engines, and other machinery 
after they are set. No drawing shall be scaled. 

The purchaser reserves the right to put other parties at work on the 
premises erecting machinery and doing other work during the con- 
tinuance of this contract. 

The contractor must conform to such rules and regulations as are 
in force on the property and carry out this contract with as little 
interference as possible with the other work of the purchaser. 

Drawings.* 

The following is a list of the drawings which accompany this 
specification and which form a part thereof: 

(Title) No. (Drawing Number) 

do do etc. 

Description of Plant. 

Steam Piping. — The plant consists of four (4) 250-horse-power 
boilers from which steam will be led to two (2) independent 16-inch 
steam headers. Two (2) 8-inch take-offs will be carried from each 
boiler, one (1) to each header. These 8-inch take-offs will consist of 
large radius wrought-iron bends of a quality hereinafter described. 
From each header a 6-inch supply will be led to the low-pressure 
cylinders of these engines. From the header nearest the engine room 
two (2) 6-inch supplies will be led through the partition wall into the 
* Drawings have been omitted. 



762 STEAM POWER PLANT ENGINEERING 

cellar for the purpose of furnishing steam to the condenser pumps. 
From one (1) header a 6-inch and 5-inch supply will be carried to a 
tandem compound engine and thence 5-inch to the two (2) exciter 
engines. 

The condenser steam lines will be continued under the engine-room 
floor so as to form a reserve steam supply to the exciter engines. 
From the end of each 16-inch main steam header a 3-inch loop will be 
carried to and from the fire and boiler feed pumps. The general 
arrangement with sizes of pipes, separators, valves, etc., is shown on 
drawing No. . 

The contractor's high-pressure steam work will begin with the steam 
nozzles on the boilers and end with the throttle valves of pumps and 
engines. The contractor will furnish all pump throttle valves but not 
engine throttle valves. 

Condenser and Exhaust Piping. — The three (3) jet condensers and 
air pumps will be set under engine-room floor. Injection water will 
be led to the condenser through a 16-inch cast-iron pipe which begins 
at a point 6 feet outside the wall of the boiler house and runs from 
this point to the condensers. After leaving the connections for the 
fourth condenser, this main reduces to 14 inches in diameter; after 
leaving the third it reduces to 10 inches in diameter; and after 
leaving the second it reduces to 3 inches in diameter. 

This injection water main will be connected with the city supply by 
a 5-inch cast-iron water pipe; and the discharge pipe will be of cast- 
iron and will be led from the condensers, beginning with a diameter of 
5 inches, increasing to 10 inches, and again increasing to 14 inches, 
and then to 16 inches after the last condenser, from whence it will be 
carried to a point 6 feet outside of the boiler-house wall. 

The low-pressure cylinder of the tandem compound engine will be 
connected by a 10-inch and a 12-inch pipe with No. 1 condenser. 
Connection will be made between both high- and low-pressure cylin- 
ders of the 750-horse-power engines with 18-inch condenser exhaust 
headers. Between the exhaust header and the cylinder of the engines, 
connections will be made with a free exhaust main leading to and 
from exhaust riser extending through the roof of. the boiler house. 
The general arrangement and the sizes of the gate valves, free exhaust 
valves, etc., are shown on drawing No. . 

The exhaust from the two (2) exciter engines and the three (3) 
condensers, the boiler feed pumps, and the fire pumps will be collected 
and led to a 1500-horse-power open-type feed- water heater located in 
the boiler room. This heater will be provided with an exhaust riser 
and exhaust head extending through the roof of the boiler house. 



TYPICAL SPECIFICATIONS 763 

The arrangement of the auxiliary exhaust pipe is shown on drawing 
No. . 

Boiler Feed Water Piping. — To the feed-water heater above 
mentioned a 2J-inch supply of fresh water will be furnished. The 
boiler feed pumps will draw the feed water from the heater through a 
6-inch cast-iron suction pipe. They will deliver it through 3-inch 
brass pipes to two (2) hot- water meters equipped with by-passes and 
thence through a 3-inch brass pipe line to boilers. A reserve wrought- 
iron 3-inch feed pipe line will be run parallel to the brass feed pipe 
line and so connected up to it that it may be used as a reserve in case 
of accident to the brass line. Connections from 3-inch lines to boilers 
will be 2J-inch brass pipe. 

The general arrangement of the feed pipe, including gate valves, 
check valves, sizes, etc., is shown on drawing No. . 

Drip Piping. — There will be two (2) systems of drip piping, 
including traps, trap by-passes, valves, etc. 

The high-pressure system comprises all the drips from the main 
steam header, separators, and high-pressure steam lines. The water 
from these traps will be led to the feed-water heater, drawing 
No. . 

The low-pressure system comprises all the drips from exhaust mains, 
receivers, and cylinder drain cocks. 

This water will be led to a catch-basin from whence it will be 
pumped by a tank pump. This pump will be furnished by the pur- 
chaser, drawing No. . 

Blow-off Pipes. — A separate and distinct wrought-iron, 2|-inch 
blow-off pipe will be run from the blow-off valve of each of the 
four (4) water-tube boilers. These pipes will be run in trenches in 
the boiler-room floor to catch-basins located outside of building. The 
purchaser will construct the trenches and catch-basins. 

Water Piping. — The purchaser will make a 6-inch connection from 
the city main to the northwest corner of the boiler room. From this 
point the contractor will run 6-inch cast-iron pipe to the inlet of 
the fire pump. From this he will run 6-inch pipe to the several fire 
hydrants. He will connect this 6-inch line with the injection water 
main with a 5-inch cast-iron connection. Near the fire pump he will 
run a 24-inch connection to the feed-water heater. The pipe around 
the fire pump will be so arranged that ordinarily city pressure will be 
maintained on the water system, but in case of need the fire pump may 
be used to increase this pressure. 

This piping is all shown on drawing No. . 

Wr ought-Iron Pipe. — All wrought-iron pipes or steam lines to be 



764 



STEAM POWER PLANT ENGINEERING 



guaranteed full- weight lap- welded wrought-iron pipe, in accordance 
with the standard dimensions as given in the table below. But the 
16-inch headers will be 16 inches O.D., 15.04 inches I.D., and shall 
weigh 87 pounds per foot. 



Nominal Internal. 


Actual External. 


Actual Internal. 


Nominal Weight per 
Foot. 


Inches. 


Inches. 


Inches. 


Pounds. 


1 


1.315 


1.048 


1.663 


H 


1.66 


1.38 


2.444 


H 


1.9 


1.611 


2.678 


2 


2.375 


2.067 


3.609 


2* 


2.875 


2.468 


5.739 


3 


3.5 


3.067 


7.536 


3* 


4 


3.548 


9.001 


4 


4.5 


4.026 


10.665 


4* 


5 


4.508 


12.34 


5 


5.563 


5.045 


14.502 


6 


6.624 


6.065 


18.762 


8 


8.625 


7.982 


28.35 


10 


10.75 


10.019 


40.065 


14 


15 


14.25 


57.893 


15 


16 


15.25 


71.77 



All pipes carrying hot water, such as feed-water, blow-offs, and 
high-pressure drip pipes, are to be extra heavy. 

All lap-welded pipe shall be proved to 500 pounds pressure per 
square inch before shipment. Butt- welded pipe shall be proved to 
300 pounds pressure per square inch before shipment. 

Exhaust pipes are to be wrought iron or steel in every case; sizes 
8 inches and above may be light O.D. tubing with flanges peened 
or expanded on, but must be absolutely tight at 28 inches vacuum. 
These pipes must be tested to 25 pounds hydrostatic pressure after 
erection. 

All high-pressure wrought-iron pipe when in place is to be tested at 
250 pounds air pressure, and must be good for a working pressure of 
175 pounds. All high-pressure steam pipe to be tested after erection 
at 175 pounds steam pressure, to the satisfaction of the engineer in 
every particular. 

Cast-Iron Pipe. — The cast-iron pipe for injection and discharge 
water and for all other purposes, as shown on the accompanying 
drawings, shall be made of tough gray iron not less than f inch thick 
at any point, free from blowholes, true to pattern, and of workman- 
like finish. It shall be tested at 250 pounds pressure before erection, 
and when in place shall be tested to 25 pounds hydrostatic pressure. 

Cast-iron pipes inside of buildings are to be flanged with flanged 
fittings. 



TYPICAL SPECIFICATIONS 



765 



Outside of buildings they shall be bell and spigot ends, and must be 
coated with tar both inside and outside. 

Valves. — All steam valves over two inches, except throttle valves, 
will be gate valves. All gate valves will be Chapman's best make or 
equivalent. 

On high-pressure steam lines, valves must be fitted with bronze re- 
movable seats, outside screw and yoke, and by-passed from 5 inches up. 

On exhaust lines, standard pattern gate valves will be used (Chap- 
man or equivalent), with inside screw. 

Globe valves on high-pressure steam and water lines will be Schutte's 
make or equivalent. 

Free exhaust valves will be either Schutte's make or equivalent. 

Exhaust valves between engines and condensers, injection water 
valves at condensers, as well as priming water valves, will have their 
stems extended and fitted with floor stands, so that they can be 
operated from floor above. Floor stands will be polished all over, 
with polished hand wheels and indicators. 

All high-pressure valves will be tested to the satisfaction of the 
engineer, at a hydrostatic pressure of 350 pounds per square inch 
and air pressure of 100 pounds between gates. 

Standard valves for exhaust and water will be tested to 150 pounds. 

Flanges and Fittings. — All flanges and fittings for high-pressure 
steam lines 3 inches and over will be open-hearth steel castings. 
These castings must be perfectly solid, made from heavy patterns, and 
free from blowholes or other defects. They must be tested at works 
to 500 pounds hydrostatic pressure, and guaranteed for a working 
steam pressure of 175 pounds per square inch. 

The diameter, thickness, and drilling of flanges must not be less than 
the dimensions given in the following table. 





Diameter of 




Number of 




Size of Pipe. 


Flange. 


Thickness. 


Bolts. 


Size of Bolts. 


Inches. 


Inches. 


Inches. 




Inch. 


16 


25 


If 


20 


1 


14 


23 


If 


16 


1 


12 


20 


14 


16 


1 


10 


m 


if 


12 


1 


9 


16 


1A 


12 


7 
8 


8 


15 


H 


12 


I 


7 


14 


1$ 


12 


3 


6 


13 


ifk 


8 


1 


5 


11 


1* 


8 


1 


H 


10* 


i& 


8 


1 

4 


4 


10 


i^ 


8 


3 

4 


3* 


9 


l 


8 


f 


3 


9 


l 


8 


f 



766 . STEAM POWER PLANT ENGINEERING 

Below 3 inches, on all high-pressure steam, water, or drip lines, extra 
heavy screwed cast-iron fittings are to be used, with sufficient extra 
heavy flange unions, so that any section of pipe can be readily taken 
out without disturbing the balance. 

The brass boiler feed piping will be made up with extra heavy brass 
screwed fittings made from cast-iron patterns. 

Pipe will be iron pipe size. 

Brass flange unions will be made from standard cast-iron patterns, 
and a sufficient number used to make up the sections readily. 

On low-pressure piping, standard fittings and flanges are to be used. 
Sizes 6 inches and above are to be flanged, and below 6 inches, screwed, 
w T ith sufficient number of flange unions in same. 

All high-pressure flanges are to be recessed on face, pipe is to be 
screwed, then peened up tight, and must stand the tests given 
above. 

Standard flanges must stand the tests as stated. 

All bolt holes must be drilled in solid metal; no cored holes will be 
allowed. 

On blow-off lines, long sweep fittings are to be used, with extra 
heavy flange unions. 

Gaskets and Bolts. — All gaskets on high-pressure steam lines must 
be copper. 

On exhaust and water lines metallic gaskets are to be used. 

All bolts must have hexagonal heads and nuts, no matter what size, 
points to be finished and nuts to be cold punched. 

Pipe Covering. — All high-pressure steam lines, separators, and all 
drip lines between high-pressure steam lines and their traps will be 
covered with an approved magnesia pipe covering. All fittings will 
have molded magnesia coverings to fit them neatly. 

Exhaust Heads. — The contractor will furnish and erect two (2) 
exhaust heads, one on the free exhaust from the engines and the other 
on the free exhaust from the feed-water heater. 

Pipe Supports. — There will be furnished and erected the necessary 
wrought-iron hangers to support the two (2) 16-inch headers in the 
boiler room. There will also be supplied the necessary supports for 
the steam mains to the pumps and exciter engines. The injection 
water main and the condenser exhaust header will be hung from the 
engine room floor beams. The free exhaust header and the discharge 
water main will be supported on piers and saddles from the cellar 
floor. The saddles to be furnished by the contractor. 

Separators. — The contractor will furnish and erect, as shown on 
drawing No. , eleven (11) steam separators. 






TYPICAL SPECIFICATIONS 



767 



Anchors. — All live-steam mains, and especially the 16-inch steam 
headers in the boiler room, will be anchored so as to secure no move- 
ment (at point of anchorage) on account of the expansion and con- 
traction or vibration. Anchors to be placed where needed after the 
rest of the work is completed. 

Relief Valves. — The boiler-feed and fire pumps shall be fitted 
with automatic controlling devices so that a certain definite maxi- 
mum pressure may be maintained on the feed-water and fire 
systems. 

Overflow Pipes. — The contractor will furnish an overflow relief valve 
for the fire pump and will connect same with the drain outside the boiler 
house. He will also connect the blow-off and overflow from the feed- 
water heater with the said drain. 

Long Radius Bends. — Where shown on the drawings, changes in 
direction of pipe runs will be made with long radius bends. These 
bends will be of wrought-iron of the quality described above. Nine 
inches of straight pipe will be left on the ends of the bends to cut 
necessary threads. 

The following table shows the minimum radii for these wrought-iron 
bends. 



Size. 


Minimum Radius. 


Size. 


Maximum Radius. 


Inches. 


Inches. 


Inches. 


Feet. Inches. 


2 


4* 


7 


3 


n 


6 


8 


3 4 


3 


8 


9 


4 


H 


10 


10 


4 4 


4 


14 


12 


5 6 


H 


16 


14 


7 


5 


20 


16 


7 6 


6 


24 













I -Beam Supports. — The contractor will furnish and erect on the 
top chord of the roof trusses in the boiler room a sufficient length of 
I beams or channels from which he will hang the 16-inch headers. 
The steam connections between 16-inch header and the engines will be 
supported by underneath rod trusses, and vibration will be taken up 
by means of lateral ties. 

Flange Covering. — After all the other work is completed and the 
plant has been in operation not less than two (2) weeks, the con- 
tractor will cover all flanges and other joints with an approved 
magnesia covering. 



768 STEAM POWER PLANT ENGINEERING 

Reducing and Relief Valves. — The contractor will furnish and erect 
two (2) reducing valves on the throttle valves of the low-pressure 
cylinders of the 750-horse-power engines. These valves will be of 
the Kiely type or equivalent. 

There will also be furnished and erected the necessary traps from 
the receivers between the high and low-pressure cylinders. 

Check Valves. — All drip lines will be equipped with check valves 
which will be set at the lowest point in the line. Two (2) three-inch 
brass checks will be placed on the boiler feed pump outlets, all as 
shown on drawing No. . 

The outlet of the boiler feed pumps will also have a relief valve 
which will return the water to the feed-water heater in case the pres- 
sure on the pump outlet should rise above a certain definite point. 

Traps. — All traps on high-pressure drip lines will be extra heavy 
steam traps. 

Painting. — All pipes shall be painted with two (2) coats of slate 
graphite and linseed oil, or other good pipe paint satisfactory to the 
engineer. 

The exhaust and condenser piping must be thoroughly painted twice 
under vacuum. 

Damper Regulator. — The purchaser will furnish the damper regu- 
lator, but the contractor will set it and connect it with both steam 
headers, water supply, and damper. 

Shields. — Where steam mains pass through the partition wall of 
the building the contractor will neatly close the opening with a sheet- 
metal shield so as to prevent dust from the boiler room from entering 
the engine room. 

Meters. — The contractor will furnish and connect as shown on 

drawing No. two (2) 3-inch hot- water meters, approved by the 

engineer, to have a capacity of 200 gallons per minute. 

Gauges. — The purchaser will furnish the customary gauges and 
boards, which will be set up and connected by the contractor. 

Machinery. — The purchaser will furnish all engines, pumps, con- 
densers, feed-water heater, boilers, foundations, but the contractor 
shall connect up the above machinery according to the evident intent 
and meaning of this specification. 

The purchaser will not furnish any pipe, valves, fittings, pipe cover- 
ing, etc., or anything connected with the work except the machinery 
mentioned above. 



TYPICAL SPECIFICATIONS 769 

416. Government Specification and Proposal for Supplying Coal. 

U. S. Treasury Department. 
United States 



PROPOSAL. 

1 Sealed proposals will be received at this office until 2 o'clock p. m., 

2 , 190 . . , for supplying coal to the United States 

3 building at 

4 as follows : 

5 : 

6 

7 '. 

8 The quantity of coal stated above is based upon the previous annual 

9 consumption, and proposals must be made upon the basis of a delivery of 

10 10 per cent more or less than this amount, subject to the actual requirements 

11 of the service 

12 Proposals must be made on this form, and include all expenses incident 

13 to the delivery and stowage of the coal, which must be delivered in such 

14 quantities, and at such times within the fiscal year ending June 30, 190 , 

15 as may be required. 

16 Proposals must be accompanied by a deposit (certified check, when 

17 practicable, in favor of ) 

18 amounting to 10 per cent of the aggregate amount of the bid submitted, as 

19 a guaranty that it is bona fide. Deposits will be returned to unsuccessful 

20 bidders immediately after award has been made, but the deposit of the 

21 successful bidder will be retained until after the coal shall have been 

22 delivered, and final settlement made therefor, as security for the faithful 

23 performance of the terms of the contract, with the understanding that the 

24 whole or a part thereof may be used to liquidate the value of any deficiencies 

25 in quality or delivery that may arise under the terms of the contract. 

26 When the amount of the contract exceeds $10,000, a bond may be exe- 

27 cuted in the sum of 25 per cent of the contract amount, and in this case, the 

28 deposit or certified check submitted with the proposal will be returned after 

29 approval of the bond. 

30 The bids will be opened in the presence of the bidders, their representa- 

31 tives, or such of them as may attend, at the time and place above specified. 

32 In determining the award of the contract, consideration will be given to 

33 the quality of the coal offered by the bidder, as well as the price per ton, 

34 and should it appear to be to the best interests of the Government to 

35 award the contract for supplying coal at a price higher than that named in 

36 lower bid or bids received, the award will be so made. 

37 The right to reject any or all bids and to waive defects is expressly 

38 reserved by the Government. 



770 STEAM POWER PLANT ENGINEERING 



DESCRIPTION OF COAL DESIRED.* 

39 Bids are desired on coal described as follows: 

40 

41 

42 

43 

44 

45 

46 : 

47 

48 

49 

50 Coals containing more than the following percentages, based upon dry 

51 coal, will not be considered: 

52 Ash per cent. 

53 Volatile matter per cent. 

54 Sulphur per cent. 

55 t Dust and fine coal as delivered at point of consumption per cent. 

DELIVERY. 

56 The coal shall be delivered in such quantities and at such tinges as the 

57 Government may direct. 

58 In this connection, it may be stated that all the available storage capacity 

59 of the coal bunkers will be placed at the disposal of the contractor to 

60 facilitate delivery of coal under favorable conditions. 

61 After verbal or written notice has been given to deliver coal under this 

62 contract, a further notice may be served in writing upon the contractor to 

63 make delivery of the coal so ordered within twenty-four hours after receipt 

64 of said second notice. 

65 Should the contractor, for any reason, fail to comply with the second 

66 request the Government will be at liberty to buy coal in the open market, 

67 and to charge against the contractor any excess in price of coal so purchased 

68 over the contract price. 

SAMPLING. 

69 Samples of the coal delivered will be taken by a • representative of the 

70 Government. 

71 In all cases where it is practicable, the coal will be sampled at the time 

* Note. — This information will be given by the Government as may be deter- 
mined by boiler and furnace equipment, operating conditions, and the local market. 

t Note. — All coal which will pass through a £-inch round-hole screen. 



TYPICAL SPECIFICATIONS 771 

72 it is being delivered to the building. In case of small deliveries, it may be 

73 necessary to take these samples from the yards or bins. The sample 

74 taken will in no case be less than the total of one hundred (100) pounds, to 

75 be selected proportionally from the lumps and fine coal, in order that it 

76 will in every respect truly represent the quality of coal under considera- 

77 tion. 

78 In order to minimize the loss in the original moisture content the gross 

79 sample will be pulverized as rapidly as possible until none of the fragments 

80 exceed | inch in diameter. The fine coal will then be mixed thoroughly 

81 and divided into four equal parts. Opposite quarters will be thrown out, 

82 and the remaining portions thoroughly mixed and again quartered, throw- 

83 ing out opposite quarters as before. This process will be continued as 

84 rapidly as possible until the final sample is reduced to such amount that all 

85 of the final sample thus obtained will be contained in the shipping can or 

86 jar and sealed air-tight. 

87 The sample will then be forwarded to the Chief Clerk of the Treasury 

88 Department, care of the storekeeper. 

89 If desired by the coal contractor, permission will be given to him, or his 

90 representative, to be present and witness the quartering and preparation of 

91 the final sample to be forwarded to the Government laboratories. 

92 Immediately on receipt of the sample, it will be analyzed and tested by 

93 the Government, following the method adopted by the American Chemical 

94 Society, and using a bomb calorimeter. A copy of the result will be mailed 

95 to the contractor upon the completion thereof. 



CAUSES FOR REJECTION. 

96 A contract entered into under the terms of this specification shall not 

97 be binding if, as the result of a practical service test of reasonable duration, 

98 the coal fails to give satisfactory results due to excessive clinkering, or to 

99 a prohibitive amount of smoke. 

100 It is understood that the coal delivered during the year will be of the 

101 same character as that specified by the contractor. It should, therefore, 

102 be supplied, as nearly as possible, from the same mine or group of mines. 

103 Coal containing percentages of volatile matter, sulphur, and dust higher 

104 than the limits indicated on line 54, and coal containing a percentage of 

105 ash in excess of the maximum limits indicated in the following table will 

106 be subject to rejection. 

107 In the case of coal which has been delivered and used for trial, or which 

108 has been consumed or remains on the premises at the time of the deter- 

109 mination of its quality, payment will be made therefor at a reduced price 

110 computed under the terms of this specification. 

111 Occasional deliveries containing ash up to the percentage indicated in 

112 the column of "Maximum limits for ash," on page 700, may be accepted. 



772 



STEAM POWER PLANT ENGINEERING 



113 Frequent or continued failure to maintain the standard established by 

114 the contractor, however, will be considered sufficient cause for cancellation 

115 of the contract. 

* PRICE AND PAYMENT. 

116 Payment will be made on the basis of the price named in the proposal 

117 for the coal specified therein, corrected for variations in heating value and 

118 ash, as shown by analysis, above and below the standard established by 

119 contractor in this proposal. For example, if the coal contains two (2) 

120 per cent, more or less, British thermal units than the established standard, 

121 the price will be increased or decreased two (2) per cent accordingly. 

122 The price will also be further corrected for the percentages of ash. For 

123 all coal which by analysis contains less ash than that established in this 

124 proposal a premium of 1 cent per ton for each whole per cent less ash will 

125 be paid. An increase in the ash content of two (2) per cent over the 

126 standard established by contractor will be tolerated without exacting a 

127 penalty for the excess of ash. When such excess exceeds two (2) per cent 

128 above the standard established, deductions will be made from price paid 

129 per ton in accordance with following table : 



Ash as estab- 
lished in 
proposal. 


No 

deduc- 
tion for 
limits 
below. 


Cents per ton to be deducted. 


Maxi- 


2 


4 


7 


12 


18 


25 


35 


mum 

limits 

for 




Per 


centages 


of ash in 


dry coal. 






ash. 


Per cent. 
5 


7 
8 
9 

10 
11 
12 
13 
14 
15 
16 
17 
18 
19 
20 


7- 8 

8- 9 
9-10 

10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 


8- 9 
9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19- 
19-20 
20-21 
21-22 


9-10 
10-11 
11-12 
12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 


10-11 
11-12 
12-13 
13-14 
14-15 
15-16 


11-12 
12-13 
13-14 


12-13 
13-14 
14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 


13-14 
14-15 
15-16 
16-17 
17-18 


12 


6 


13 


7 


14 


8 


14-15 
15-16 
16-17 
17-18 
18-19 
19-20 
20-21 
21-22 
22-23 


14 


9 


15 


10 


16 


11 


16-17 
17-18 
18-19 
19-20 
20-21 
21-22 
22-23 


16 


12 


17 


13 


18 


14 


19 


15 


19-20 
20-21 
21-22 
22-23 


19 


16 






20 


17 






21 


18 








22 















* Note. — The economic value of a fuel is affected by the actual amount of com- 
bustible matter it contains, as determined by its heating value shown in British 
thermal units per pound of fuel, and also by other factors, among which is its ash 
content. The ash content not only lowers the heating value and decreases the 
capacity of the furnace, but also materially increases the cost of handling the coal, 
the labor of firing, and the cost of the removal of ashes, etc. 



TYPICAL SPECIFICATIONS 773 

Proposals to receive consideration must be submitted upon this form and contain 
all of the information requested. 



, 190 

The undersigned hereby agree to furnish to the U. S 

building at , the coal described, in tons 

of 2,240 pounds each and in quantity, 10 per cent more or less than that stated 
on page 697, as may be required during the fiscal year ending June 30, 190 , 
in strict accordance with this specification; the coal to be delivered in such 
quantities and at such times as the Government may direct. 

Price per ton (2,240 pounds) $ 

Commercial name of the coal 

Name of the mine or mines 

Location of the mine or mines ' 

Name or other designation of the coal bed or vein 

Size (indicate information which will apply) — 

Unsized Lump Run of mine 

(Round ) ~ . 

Screened, through inch and over inch -j Square ) 

iBar screen. 
Data to establish a basis for payment: 

British thermal units in coal as delivered 

Ash in dry coal (Method of American Chemical Society) per cent. 

It is important that the above information does not establish a higher standard than can be 
actually maintained under the terms of the contact; and in this connection it should be noted 
that the small samples taken from the mine are invariably of higher quality than the coal actually 
delivered therefrom. It is evident, therefore, that it will be to the best interests of the contractor 
to furnish a correct description with average values of the coal offered, as a failure to maintain the 
standard established by contractor will result in deductions from the contract price, and may 
cause a cancellation of the contract, while deliveries of a coal of higher grade than quoted will be 
paid for at an increased price. 



Signature : . . . . 
Address : 



Name of corporation, 

Name of president, 

Name of secretary, 

Under what law (State) corporation is organized : . 



CHAPTER XX. 

A TYPICAL CENTRAL STATION. 
Steam Turbines. — Alternating Currents. 

The Fisk Street Station of the Commonwealth Edison Company, 
Chicago, is an excellent example of modern central station practice. 
The present (June, 1910) rated capacity of the plant is 120,000 kilo- 
watts, though space is available for a considerable increase. The 
station is located on the banks of the Chicago River near Fisk and 
Twenty-second streets, as indicated in Fig. 438, and is about one and 
one-half miles south of the center of distribution of the present load. 
The location of the station between the east and west slips of the river 
secures an unusual advantage in the location of the intake and discharge 
tunnels, and the extent of property affords ample storage capacity for 
coal. Both the Chicago & Alton and the Chicago, Burlington & Quincy 
railways extend into the property, giving excellent facilities for the 
transportation of coal, ashes, construction materials, and machinery. 
The plant is constructed on the unit basis, each turbine and generator 
having its own boilers, auxiliaries, and piping system, thus permitting 
any unit to be shut down without interfering with the operation of the 
rest of the system. 

Building. — The main building rests on piles, driven to hard pan, 
capped with a grillage of I beams and concrete. The walls are of red 
pressed brick trimmed with white Bedford stone. The windows are 
25 feet wide and 32 feet high, the sections of which are operated by 
compressed air. Fig. 439 gives a view of the north elevation. Large 
skylights afford ample light and ventilation. The entire interior wall 
surface of the turbine room is finished with white enameled brick 
trimmed with terra cotta. The boiler-room walls have an eight-foot 
wainscoting of enameled brick, the remainder being red pressed brick. 
The floors are of concrete, that in the turbine room being covered with 
two-inch hexagonal terra-cotta tile. The roofs are of Roebling concrete. 
The total width of the building is 243 feet, the turbine room taking 
up 61 feet, the boiler room 142 feet, and a car track the remainder. 
A 50-ton motor-driven crane spans the turbine room and is used in 
connection with units 1 to 4, and a 60-ton crane is provided for the 

774 



A TYPICAL CENTRAL STATION 



775 




776 



STEAM POWER PLANT ENGINEERING 



remaining units. A 5-ton auxiliary hoist is also provided on the main 

cranes. In the boiler room a 
small hand-power crane serves 
each two batteries of boilers. 

Coal and Ash Handling. — 
An interior shed extends the 
entire length of the east end 
of the building, as indicated 
in Figs. 440 and 441. Coal 
is brought in on cars and 
dumped or shoveled into a 
track hopper, from which it 
is delivered to the overhead 
bunkers by the conveying 
system. A crusher is placed 
between the track hopper and 
main conveyor to be used in 
case lump coal is furnished. 
These bunkers have a capacity 
of 1200 tons each, sufficient 
for several days run. The 
conveyors are driven by a 
15-horse-power motor and are 
of the McCaslin pattern, end- 
less chain, with overlapping 
buckets, each bucket having 
a capacity of 100 pounds. 
The conveyors move at a rate 
of 50 feet per minute, giving 
a service capacity of 75 tons 
per hour for each unit. A 
separate conveyor and bunker 
is installed in each section of 
16 boilers. The coal bunkers 
feed through flexible down 
spouts to the stoker maga- 
zines. Underneath the front 
end of the stoker is a fine-coal 
hopper which collects the fine 
coal falling through the grate 
and discharges it into the conveyor system, as in Fig. 99. The ashes 
collect in the ash pit, from which they are dumped into the conveyor 





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A TYPICAL CENTRAL STATION 



777 




778 



STEAM POWER PLANT ENGINEERING 




A TYPICAL CENTRAL STATION 779 

and carried to an ash bin directly over the coal track. Illinois screen- 
ings furnish the greater part of the fuel. Provision is also made for 
outside storage. 

Boilers. — The boiler plant is divided into five sections, each section 
consisting of sixteen 500-horse-power B. & W. boilers arranged in bat- 
teries of eight and equipped with B. & W. chain grates. The settings 
are installed back to back, as illustrated in Fig. 440. Each boiler has 
two 42-inch steam drums, approximately 5000 square feet of heating 
surface, and about 1000 square feet of superheating surface. Steam 
is generated at a pressure of 200 pounds per square inch with super- 
heat of 150 degrees F. The ratio of water-heating surface to grate 
surface is approximately 55 to 1, and the ratio of the total heating 
surface to grate surface is about 66 to 1. When burning Illinois screen- 
ings, an average depth of 7 inches is maintained on the grate, with speed 
of grate of 5 inches per minute. The grates are driven by Krehbiel 
oscillating engines, two engines being provided for each section but 
only one being in use at a time. The boilers are supported by reen- 
forced girders of the main building structure. A gallery is placed in 
front of the settings, 8 feet above the floor, to facilitate cleaning of 
tubes. Galleries are also placed between the batteries and on top of 
them. Spaces of 5 feet are provided between the sides and rears of the 
batteries, and 18 feet 8 inches in front. The furnaces are similar to 
the one illustrated in Fig. 62. The outside of the setting is finished 
with red pressed brick. Each drum is fitted with a four-inch pop 
safety valve. The blow-off main is 4 inches in diameter and discharges 
into the river. There are four blow-off connections to each boiler, each 
being provided with a blow-off cock and an angle valve; three of the 
connections are fitted to the mud drum and the other to the super- 
heater drain. 

Chimneys. — One stack is provided for each section of 16 boilers. 
The shaft is supported by the steel work of the boiler setting, as 
shown in Fig. 442, an arrangement which commends itself where 
space is limited and real estate values are high. The stacks for all 
units are 259 feet 6 inches in height above the grate surface, and are 
18 feet in internal diameter. The lining is of radial fire brick and 
varies from 4 inches to 13 inches in thickness. The steel sections are 
5 feet high and vary in thickness from f inch to \ inch. . There are 
two flues, one 32 feet long and the other 63 feet, which enter the stack 
on opposite sides. 

Turbines. — The prime movers are vertical five-stage Curtis turbines 
with base condenser and are rated at 12^000 kilowatts each. The normal 
speed is 750 r.p.m. The average steam consumption, including all 



80 



STEAM POWER PLANT ENGINEERING 




~^fON mooji Satiij" 



A TYPICAL CENTRAL STATION 781 

auxiliaries is approximately 15 pounds per kilowatt hour, corresponding 
to a coal consumption of 3 pounds per kilowatt hour (Illinois screenings, 
10,400 B.T.U. per pound). Special tests have shown as low as 12.8 
pounds per kilowatt hour, initial pressure 200 pounds gauge, 150 de- 
grees superheat, absolute back pressure \ inch of mercury. Each pair 
of units has a pair of duplicate pumps, an accumulator and a storage 
tank for supplying oil, the step-bearing pressure being maintained at 
750 pounds per square inch. When the accumulator falls below a 
certain point a motor-driven pump is automatically started. 

Generators. — The generators are 2300- volt, 25-cycle, three-phase 
General Electric alternators mounted over the vertical shaft as illus- 
trated in Fig. 185. Exciting current is furnished by 

2 50-kilowatt motor-driven generators. 
2 75-kilowatt motor-driven generators. 
2 1 50-kilowatt motor-driven generators. 
2 75-kilowatt steam-driven generators. 
2 1 50-kilowatt steam-driven generators. 

Part are held in reserve, though no particular units are maintained 
for the purpose. The high-tension cables lead -from the generator 
through an underground tunnel to the switch house, located about 
50 feet west of the main building. The oil switches, wattmeters and 
other instruments are located on the first floor, while the bus-bars and 
other high-tension connections are in the basement. The station switch- 
board or operating gallery in the main building is equipped with only 
such devices as are necessary for the control of the machines, all other 
instruments being located in the switch house. From the switch house 
the high-tension current is conducted through oil switches to the various 
substations, where it is converted to direct current by rotary converters, 
or transformed from 25 to 60 cycles by motor generator sets. 

The Twenty-second Street substation is located at the north end of 
the property (Fig. 438). In this substation are installed two motor 
generator sets and one rotary converter, the latter supplying direct 
current to the neighboring district and to the main station. 

Boiler and Turbine Piping. — Immediately below each boiler section 
is an apartment called the " header room," where the steam pipes from 
the various boilers join the main header, which increases in size from 
6 inches at the most remote boiler to 10 inches at the middle boiler, 
and finally to 14 inches where it leaves the nearest boiler and passes to 
the turbines. The pipes are of wrought iron, with welded flanges, and 
are packed with copper gaskets. The feeder from each boiler is 6 inches 
in diameter. An angle stop valve and a check valve are placed at the 



782 STEAM POWER PLANT ENGINEERING 

boiler nozzle and a globe valve at the header. A motor-operated throttle 
valve and strainer are placed at the turbine, and a hydraulically oper 
ated valve, controlled from the operating gallery, is located in the 
header room. The main header is not anchored at any point, the entire 
weight being carried by roller supports. The only drain in the header 
is a f-inch bleeder on the turbine side of the hydraulically operated 
valve. The bleeder is connected to a trap which discharges into either 
boiler or superheater blow-off main. All branches or feeders are drained 
and discharged into the superheater blow-off. 

Condensers and Auxiliaries. — Each unit has its own condensing 
apparatus, feed-water heater, hot well and feed pumps. The condensers 
are of the Worthington " base " type with 25,000 sq. ft. of cooling 
surface each, composed of 5900-6000 1-inch tubes 16 feet long. Cooling 
water is taken from the east slip through concrete tunnels and is dis- 
charged from the condenser into similar tunnels which empty into the 
west slip. (See Fig. 438.) 

The dry vacuum pumps are of the rotative type, with cylinders 
26 X 24, r.p.m. 100-120, and are driven by a 75-horse-power Corliss 
engine. 

The circulating pumps are of the volute single-stage centrifugal type 
and are mounted on an extension of the main shaft of the engine driving 
the dry vacuum pump. They are rated at 22,500 gallons per minute 
each. 

The hot-well pumps are of the two-stage centrifugal type, driven by 
20-horse-power direct-current motors. 

The feed pumps are of the Dean vertical single-cylinder pattern and 
are installed in duplicate for each unit. The steam cylinders are 24 
inches in diameter, water cylinders 14 inches in diameter, and common 
stroke 24 inches. Feed water is drawn in by suction from the hot well 
at a temperature of about 100 degrees F. and is forced through closed 
heaters having 3000 sq. ft. of heating surface, and its temperature is 
raised to 180 degrees. The heater receives the steam exhausted from 
the steam-driven auxiliaries. 

From the heater this water is forced through a 5-inch feed main to the 
different boilers in the section. The branches from main to boiler drum 
are 3 inches in diameter and fitted with an angle stop valve, a regrinding 
check and a grate-regulating valve. There is a 5-inch auxiliary main 
which supplies cold water to the boiler in case the main header is shut 
down. 



A TYPICAL CENTRAL STATION 



783 



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Fig. 443. General Plan of Quarry Street Station, Units 1 and 2. 



784 STEAM POWER PLANT ENGINEERING 

Miscellaneous. — The work is divided into eight-nour shifts. The 
list of operating men per unit is as follows: 

In turbine room, including janitor work 2.0 

In oil room X . 

Attending water 0.5 

Fireman 1 

Fireman's helper 1 

Conveyor men 2 

Turbine switchboard gallery 0.3 

Exciter tenders : . ■ o . 2 

Switchboard attendants „ 0.2 

Drains and drips from the auxiliaries empty into sumps from which 
they are discharged by Yeoman's bilge pumps into the discharge tunnel. 

A steam-driven house pump is located in the basement. 

The fire-protection system includes a 220-horse-power motor-driven 
spherical pump located in the basement and a connection for a fire tug. 

Dining room, reading room, shower and tub baths and sleeping 
rooms for emergencies are provided for the employees. 

Quarry Street Station. — Figs. 443 to 445 give general views of the 
Quarry Street Station, which is located directly across the river from 
the Fisk Street Station. The two stations are distinct; a breakdown 
in one would not affect the other; nevertheless, they are operated to- 
gether. That is, there is one chief engineer for the two, and the com- 
bined station force of 350 men can be shifted from one to the other as 
needed. 

The general layout of the Quarry Street Station differs from that of 
the Fisk Street Station on account of real-estate limitations. The 
boilers are in two parallel rows instead of the equipment for each unit 
extending at right angles to the turbine room as at Fisk. 

The complete station will contain six 14,000-kilowatt Curtis turbo- 
alternators, the steam for each unit being supplied by eight 500 
horse-power B. & W. boilers arranged as shown in Fig. 443. Steam is 
generated at 225 pounds gauge pressure and 150-175 degrees super- 
heat. The settings for the first two units are similar to that illus- 
trated in Fig. 67a, and the other similar to the one illustrated in Fig. 67b. 

A novel system of ventilation enables the generators to be operated 
continuously at full load. As will be seen from Fig. 444 air ducts lead 
from an outside intake to the top of each unit, the revolving portion of 
the generator being designed to draw in a continuous supply of air and 
discharge it through openings in the turbine casing. 

Everything is in duplicate, so that the chance of breakdown is remote 
There are two volute circulating pumps, each driven by a 125 horse- 



A TYPICAL CENTRAL STATION 



785 




786 



STEAM POWER PLANT ENGINEERING 



power Corliss engine. The water of condensation is removed from the 
condensers by two rotary pumps driven by Kerr steam turbines. For 
each two units of turbines and boilers three horizontal boiler feed pumps 
are provided, located between the turbines as stated. There are also 
four step-bearing oil pumps, two oil accumulators, dry-air pumps, oil 
filters, etc. All are in plain view of the turbine-room operating force. 

For the six ultimate units five 150-kilowatt exciters will be installed, 
three driven by horizontal Curtis steam turbines and two by 25-cycle, 
220- volt induction motors. One of each of these types is included in 
the equipment of the present plant. In addition there is an excitation 
storage battery of 70 cells in the basement. Furthermore, in an emer- 
gency the split-pole rotary converter of the substation could be used 
for excitation. 

Northwest Stations. — Two new stations of 120,000 kilowatt rated 
capacity each are to be installed on the north branch of the Chicago 
River near Roscoe Street and California Avenue. Each station is to be 
equipped with six 20,000-kilowatt Curtis turbo-generators, 2300 volt, 
25 cycle, three phase, 750 r.p.m., similar to those installed at Quarry 
Street. The first two units for one station are now in course of erection. 
Each unit is to be supplied with steam at 250 pounds gauge pressure and 
150 degrees superheat from ten 500-horse-power B. & W. boilers. 
The boiler settings are to be similar to the one illustrated in Fig. 67b. 
The general layout will be the same as at Fisk, the boiler lanes extend- 
ing at right angles to the turbine room. There will be one chimney 
17 feet inside diameter and 250 feet in height for every ten boilers. 

The present capacity (July, 1910) of the Commonwealth Edison 
Company is about 240,000 kilowatts, divided as follows: 





Units. 


Present 
Capacity. 


Ultimate 

Capacity. 


Fisk 


10-12,000 
6-14,000 
6-20,000 
6-20,000 


120,000 
84,000 


120,000* 


Quarry 


84,000 


Northwest No. 1 


120,000 


Northwest No. 2 




120,000 


Miscellaneous Plants 


36.000 


36,000 


m 




240,000 


480,000 



* Space is available for four additional units, but no increase is contemplated at present. 
COMPARATIVE BOILER ROOM AND ENGINE ROOM AREAS. 



Boiler room, sq. ft. per kw. 
Engine room, sq. ft. per kw 
Total area, sq. ft. per kw.. . 



Fisk. 



0.51 
0.24 
0.75 



Quarry, 



0.18 



Northwest. 



0.44 
0.15 
0.59 



A TYPICAL CENTRAL STATION 



787 




Fig. 445. Sectional Elevation, Quarry Street Station. 



CHAPTER XXI. 

A TYPICAL ISOLATED STATION. — MANUFACTURING PLANT. 

The West Albany Power Station.* 

This modern and well-designed station was erected (1904) in place 
of an old and practically worn-out plant, and is an example of what 
may be accomplished with a limited appropriation. 

The power generated is used about the repair and construction shops 
of the New York Central Railroad Company at West Albany, N.Y., 
for lighting and general power purposes, and steam is used for heating. 
The equipment includes two direct-connected 600-kilowatt generators 
and two direct-acting air compressors. 

General. — The plant (Fig. 446) is housed in a brick building 92 
feet 8 inches wide by 113 feet 4 inches long, divided by a fire wall into 
an engine and a boiler room, the former being 46 feet wide. 

The main units consist of two cross-compound non-condensing 
Ball & Wood engines and two Chicago Pneumatic Tool Company's air 
compressors. Steam is supplied by two batteries of horizontal water- 
tube boilers, comprising four boilers of 500 horse power each, made by 
the Franklin Boiler Works Company. These boilers have a guar- 
anteed normal rating of 528 horse power under a natural draft not 
exceeding t 7 q inch of water measured in the flue adjacent to the boiler, 
and an evaporation of not less than 9 pounds of water per pound of 
coal from and at 212 degrees F. They contain 5280 square feet of 
heating surface and 84 square feet of grate surface, a ratio of 63 to 1. 
There are 300 31-inch tubes 13 feet long, surmounted by two 48-inch 
steam drums. The steam drums are connected by a 5-inch balance 
pipe on which two safety valves are mounted. 

The boilers are suspended from wrought-iron and steel frames 
entirely independent of the brickwork, eliminating strains due to 
expansion and contraction. 

The smoke flues run over the boilers to the stack placed between the 
two batteries, as shown in Figs. 447 and 448. Dampers are placed in 
both flues and hoods, the former operated by automatic regulators and 
the latter by hand, although the flue dampers may be readily discon- 
nected and also operated by hand. 

* See Power, November, 1904. 

788 



A TYPICAL ISOLATED STATION 



789 



The steam nozzles are connected to dry pipes in each drum and to 
elbows outside, between which is a cross-over pipe carrying the main 
10-inch valve, which is of the automatic, non-return, hand and 
emergency angle type. Between the angle stop valve and the main 
near the elbow is a gate valve. Long bends in the 10-inch pipe 
between stop valve and main are relied on for the necessary flexibility, 
and the main header is anchored to the stack near its middle point, as 
seen in Fig. 450, and its ends allowed free movement. 




A specially designed roller suspension is used on the larger pipes, and 
where pipes are swung from the floor beams a turn-buckle is provided 
in each suspension rod for adjustment. 



790 



STEAM POWER PLANT ENGINEERING 




m"-~ 



A TYPICAL ISOLATED STATION 



791' 




792 STEAM POWER PLANT ENGINEERING 

The steam piping slopes so as to drain to the drop legs, the drips 
being returned to the boiler by the Holly return system. Steam is 
led to the two large engines through 9-inch pipes, as shown in 
Figs. 449 and 450, a separator being placed just above each engine 
throttle. 

Engines. — The engines have cylinders 21 and 41 inches in diameter 
and a stroke of 30 inches. They develop the normal load of 900 
horse power, cutting off at 37 per cent of the stroke in each cylinder 
with steam at 175 pounds pressure and running at 120 revolutions per 
minute. The guaranteed steam consumption at this load is 19 pounds 
per indicated horse power. The generators, made by the General 
Electric Company, are mounted between the cylinders and supply 
three-phase alternating current, 60 cycles per second at 480 volts. 

Speed regulation is sufficiently close to allow parallel operation of 
the generators, so that current is delivered to one bus-bar on the main 
switchboard in the engine room. Exhaust steam from the high-pres- 
sure cylinder is led to a vertical receiver below the engine-room floor, 
as shown in Figs. 449 and 450, and from there to the low-pressure 
cylinder. The low-pressure cylinders exhaust through 16-inch pipes 
into a 24-inch main extending nearly the whole length of the power 
house. This main also receives the exhausts from the exciter units, 
air compressors, and auxiliaries. 

Exciters. — The two steam-driven exciter units consist of 35-kilo- 
watt dynamos direct connected to 60-horse-power Woodbury engines 
made by the A. D. Granger Company. The engines are 7x12 single- 
cylinder automatic, and at 300 revolutions per minute and a normal 
cut-off of 20 per cent develop 58 horse power. 

Besides these steam-driven units is a 100-kilowatt General Electric 
motor-generator. Fig. 451 gives a diagrammatic outline of the switch- 
board connections. 

The two air compressors were made by the Chicago Pneumatic 
Tool Company and are of the cross-compound, two-stage type. The 
steam cylinders are 16 and 27 inches in diameter, the air cylinders are 
14 and 24 inches in diameter, the common stroke 18 inches, and the 
capacity 1225 cubic feet of free air per minute. 

Steam Piping. — Steam is supplied to the exciter engines and air 
compressors through an auxiliary steam line which runs nearly the 
length of the power house in the basement beneath the engine room 
as shown in Figs. 449 and 450. It is anchored near each end, and the 
long radius vertical bend near its middle point allows for expansion. 
It draws steam from the main steam pipe in the boiler room through 
the vertical pipe shown in Figs. 446 and 447. This pipe is taken out of 



794 



STEAM POWER PLANT ENGINEERING 



the main at the top, makes a semicircular bend and passes through 
the wall between the engine and boiler rooms and enters the auxiliary- 
main at the top. 

Connections are made from the auxiliary main to the low-pressure 
cylinders of the 900-horse-power compound engines so that they may 




be supplied with high-pressure steam should an occasion arise which 
would demand it. 

The use of an auxiliary main may at first seem extravagant, but a 
careful study of the piping arrangement will show that it does not 
entail the use of more pipe or fittings and is a very desirable arrange- 



A TYPICAL ISOLATED STATION 



795 




796 STEAM POWER PLANT ENGINEERING 

ment, particularly since it leaves the lines to the main engines intact 
and direct. 

Connections to the other engines are direct, practically straight, and 
simple, and it is easy to make connections in both the steam and 
exhaust lines with any other apparatus which may be added in the 
future. In fact, the simplicity and yet completeness of the piping 
scheme is perhaps one feature of the plant. 

Connection is made at the left of the auxiliary steam line, as shown 
in Figs. 449 and 450, to a steam supply for the shops at high pressure, 
or through a reducing valve at low pressure for heating. 

The exhaust line, which runs nearly the length of the building in the 
basement and receives exhaust steam from all sources, shunts part of 
it through an exhaust muffler and oil separator to the shops for 
heating and the rest of it through an open heater and out of the 
exhaust pipe through the roof (Figs. 448 to 450). Either the muffler 
or the heater may be by-passed and the exhaust caused to flow 
directly to the heating system or the atmosphere. 

Expansion in the exhaust line is provided for by two copper expansion 
joints, and an approved exhaust head is placed on the line extending 
above the roof. 

All high-pressure steam and water piping is of " special full weight " 
lap- welded pipe and all low-pressure pipe of " standard " weight. Steam 
and hot-water pipes are covered with the best non-conducting cover- 
ing. The high-pressure joints are made up with Merworth copper 
gaskets. 

Two duplex Worthington feed pumps, made by the International 
Steam Pump Company, supply the boiler feed through the lines shown 
in Figs. 446, 447, and 450. The source of water supply is the city 
mains; hence the use of the meters. 

Figs. 447 and 450 show the arrangement of the blow-off piping. 
There is a system of piping for fire protection which is connected to the 
city supply and runs in gradually reduced sizes as shown in Fig. 450. 
This system is laid out so as to drain itself through an arrangement of 
pipes discharging into the sewer. 

Chimney. — A sectional view of the stack which furnishes draft for 
all of the boilers is shown in Fig. 446. It is 10 feet in diameter on the 
inside and 165 feet high, and designed to provide a draft of 50 per cent 
above the normal rating of the boilers, equivalent to 1.1 inches of 
water with the temperature of the heated gases not over 500 degrees F. 
It is made of radial bricks and has a baffle wall through the center 
for a short distance from the base to prevent the gases from the flues 
on each side from impinging and causing eddies. The wall is 20J inches 



A TYPICAL ISOLATED STATION 797 

thick at the base and 7J at the top. The stack was built by M. W. 
Kellogg & Co. 

The appropriation for the power plant did not allow the installation 
of a complete coal and ash-handling apparatus, but provision was 
made for such an installation in the future, and for the present the 
apparatus shown in Fig. 448 is used. The ashes fall from the grate 
to the ashpit (Fig. 447), and are raked out through the door into a 
barrow in the adjacent passageway, from which they are discharged 
into the vertical conveyor (Fig. 448) and transferred to a hopper and 
chute suspended from the roof beams and from there through a spout 
to cars outside. The coal is delivered over this track into the coal bin 
and carried to the boiler fronts in barrows. It is seen that the 
arrangement allows for the installation of a complete coal and ash- 
handling apparatus of the continuous type, and indeed one of the 
features of design is the provision for the installation of modern appli- 
ances in the future which are not possible at present. 

The general arrangement of the boilers, engines, and piping is such 
that additional units of each may be installed by enlarging the building 
at one end without disarrangement of the apparatus already in place. 



APPENDIX A. 

GENERAL BIBLIOGRAPHY — POWER PLANT ENGINEERING. 
DESCRIPTION OF POWER PLANTS. 

Gas Driven. Central Stations. 

A Power Plant with 500-Horse-Power Gas Engines, 

Engineering Record 51 : 178 Feb. 18, 1905 

Gas and Electric Power Plant for a Railroad 

Terminal, Engineer (United States) 42: 233 April 1, 1905 

Gas-Driven Electric Power Station at Bridgewater, 

England, Power 25: 273 May, 1905 

Power Installation at Isle of Elba for Blast Furnace 

Gas, Power 24: 460, 549 

Gas Power Plant at Haysam Harbor, Power 25: 212 April, 1905 

Lackawanna Steel Company's Power Plant, Power 23:663 Dec, 1903 

Gas Power Plant for the Morecambe Bay Harbor 

Works, Engineering Record 50: 467 Oct. 15, 1904 

Power and Mining Machinery Company's Gas and 

Electric Plant, Electrical World 44: 442 Sept. 10, 1904 

1020 Dec. 10, 1904 

Engineering Record 50: 652 Dec. 3, 1904 

Engineer (United States) 41 : 829 Dec. 15, 1904 

Some Features of the Warren Gas Power Plant, 

Electrical Journal 3: 205 April, 1906 

Proceedings of Engineers' Society of Western 

Pennsylvania 22: 290 July, 1906 

Trials of Suction Gas Producer Plants, Mechanical 

Engineering 16: 707 Nov. 11, 1905 

Power 26: 287 May, 1906 

Engineering News 55: 538 May 17, 1906 

Engineering 82: 205 Aug. 10, 1906 

A Year's Experience with Gas Engines, Power 27: 831 Dec, 1907 

The Producer Gas Plant, Railway and Engineering 

Review 47:8 Jan. 5, 1907 

Engineer (United States) 44 Aug. 15, 1907 

Electrician (London) 58: 642 Feb. 8, 1907 

Electrical Review (London) 60: 201 Feb. 1, 1907 

Journal Association of Engineering Societies. . . 38: 14 Jan., 1907 

Central Station June, 1907 

Hydraulic. 

Animas Power Development in Colorado, Engineer- 
ing Record 50: 519 Oct. 29, 1904 

Engineering News 50: 616 June 3, 1905 

55: 1 Jan. 4, 1906 

798 



APPENDIX A 



799 



Hydraulic — Continued. 

American River Electric Company's System, 

American Electrician 17:1 Jan., 1905 

Water Power on the Apple River, Engineering News 54: 374 Oct. 12, 1905 

Engineering Record 52: 431 Oct. 14, 1905 

Atlanta, Georgia, Engineering Record 49: 504 April 23, 1904 

Berkshire Hydro-Electric Plant (H. S. Knowlton), 

Electrical Age 36: 107 Feb., 1906 

Bay Counties Power Company, California (I. D. 

Galloway), Engineering News 46: 230 Oct. 3, 1901 

British Columbia Mines Power and Light, Western 

Electrician 32: 321 April 25, 1903 

Catawba River Power Development near Rock 

Hill, South Carolina, Engineering Record 50: 114 July 23, 1904 

129 July 3d, 1904 

Chelton Hills, Pennsylvania, Electrical World 35: 487 March 31, 1900 

Champ Generating Station, American Electrician ... 17: 65 Feb., 1905 

Chicago Drainage Canal, Engineering News 55: 52 Jan. 18, 1906 

Columbus Power Company, Georgia, Engineering 

Record 49: 64 Jan. 16, 1904 

Chattanooga and Tennessee Power Company, 

Engineering Record 52: 576 Nov. 4, 1905 

Chittenden Power Company, Rutland, Vermont, 

Engineering Record 52: 653 Dec. 9, 1905 

Cornell University Power Plant, Engineering Record 51 : 562 May 20, 1905 
Central California Electric Company, Engineer 

(United States) 40: 160 Feb. 16, 1903 

Cudahy Packing Company, Engineer (United States) 38: 82 March 1, 1901 

Western Electrician 28: 297 May 4, 1901 

Central Lard Company, Engineering Record 46: 122 Aug. 9, 1902 

Carpenter Steel Works, Engineering Record 46: 220 Sept. 6, 1902 

Deere and Company, Moline, Illinois, Engineer 

(United States) 40: 617 Aug. 15, 1903 

Deering Harvester Company, American Electrician 12: 381 Aug., 1900 

Dixon Crucible Company, American Electrician. ... 13: 517 Nov., 1901 
Edison Portland Cement Company, Electrical 

World 42: 1051 Dec. 26, 1903 

Ford Plate Glass Company, Engineer (United 

States) 38: 107 April 1, 1901 

Henry Heide Candy Company, Engineering Record 51 : 269 May 20, 1905 
Ingersoll-Sargent Drill Company, American 

Electrician 17: 237 May, 1905 

Power 25: 323 June, 1905 

Littleton Creamery, Denver, Engineering Record . . . 50: 159 Aug. 6, 1904 
Laray Mills, Gastonia, North Carolina, Engineer 

(United States) 40: 222 March 16, 1903 

Ladew Factory, Power 25: 468 Aug., 1905 

Patton Paint Company, Newark, New Jersey, 

Electrical World. 42: 454 Sept. 12, 1903 

Engineer (United States) 40: 798 Oct. 15, 1903 



800 STEAM POWER PLANT ENGINEERING 

Hydraulic — Continued. 

Modern Malt House, Western Electrician 37: 101 Aug. 5, 1905 

John Mehl Factory, Jersey City, Engineer (United 

States) 38: 245 July 1, 1901 

Manchester, New Hampshire, Mills, Electrical 

World 41:269 Feb. 14, 1903 

Monarch Mills, Union, South Carolina, Engineering 

Record 50: 380 Sept. 24, 1904 

Lake Superior Company, St. Mary's Falls, Michigan, 

Engineering News 40: 68 Aug., 1898 

Montmorency Falls, Canada, American Elec- 
trician 12: 553 Dec, 1900 

Missouri River Power Company, American Elec- 
trician 14: 323 July, 1902 

Montreal Power Plant, Engineer (London) 101 : 130 Feb. 9, 1906 

Monterey Gas and Electric Company, Engineer 

(United States) 42: 87 Jan. 16, 1905 

Montgomery, Alabama, Engineering News 46: 418 Dec. 5, 1901 

Massena, New York, Power Plant, Engineering 

News 45:130 Feb. 21, 1901 

Engineering Record 41 : 2 Jan. 6, 1900 

122 Feb. 10, 1900 

Western Electrician 22 : 22 Jan. 8, 1898 

Mexican Light and Power Company, Engineering 

Record 51: 575 May 20, 1905 

Power Plant of the Milano-Gallarate-Porto Ceresio 

Railway, Engineer (United States) 40: 25 Jan. 1, 1903 

American Electrician 14: 556 Dec, 1902 

New Water Power Transmission Plants (C. L. 

Fitch), Cassier's 19: 243 Feb., 1901 

North Mountain Power Company, Engineer 

(United States) 43: 123 Feb. 1, 1906 

Engineering Record 53 : 27 Jan. 6, 1906 

NewMilford, Connecticut, Engineering Record 49: 187 Feb. 13, 1904 

230 Feb. 20, 1904 
Northern California Power Company, Engineering 

Record 50: 506 Oct. 29, 1904 

Niagara (J. E. Woodbridge), American Electrician. 12:1 Jan., 1900 
Transactions American Society of Mechanical 

Engineers 19: 839 June, 1898 

Extension of Niagara Falls Power Plant, American 

Machinist 21 : 463 June 23, 1898 

Western Electrician 22: 137 March 5, 1898 

Niagara Falls Power (H. W. Buck), Cassier's 20: 1 May, 1901 

179 Jan., 1902 

25: 103 Dec, 1903 

Engineering 74: 637 Nov. 12, 1902 

Engineering News 48: 9 July 3, 1902 

250 Oct. 2, 1902 

490 Dec. 11, 1902 



APPENDIX A 801 

Hydraulic — Continued. 

Power Plant of the Niagara Falls (Coleman Sellers), 

Engineering 67: 91 Jan. 20, 1899 

128 Jan. 27, 1899 

160 Feb. 3, 1899 
Canadian District Plant (C. B. Smith), Engineering 

Magazine 28: 727 Feb., 1905 

New Wheel Pit Niagara Falls Power Company 

(O. E. Dunlap), Engineering News 43: 229 April 5, 1900 

Engineering Record 43: 150 Feb. 16, 1901 

Power of Lower Niagara (O. E. Dunlap), Western 

Electrician 32: 360 June 18, 1898 

Niagara Falls (L. B. Stillwell), Western Elec- 
trician 29: 191 Sept. 21, 1900 

Progress on Power House at Niagara (O. E. 

Dunlap), Western Electrician 29: 234 Oct. 12, 1901 

Niagara Power Plant of Ontario, Engineering News 54: 475 Nov. 9, 1905 

561 Nov. 30, 1905 
Power Station at Neuchatel, American Elec- 
trician 16: 493 Oct., 1904 

Norwegian Hydro-Electric Plant, Electrical World . . 46 : 135 July 22, 1905 

Western Electrician 37: 63 July 22, 1905 

Hydro-Electric Power Plant of Christiania, Norway, 

Engineer (United States) 40: 227 March 16, 1903 

Works of the Ontario Power Company, Engineering 

Record 50: 420 Oct. 8, 1904 

460 Oct. 15, 1904 

480 Oct. 22, 1904 

29 Oct. 29, 1904 

Oliver Plow Company, Engineer (United States) ... 42: 363 June 1, 1905 

Ontario Power Company, Engineering Record 50: 420 Oct. 8, 1904 

460 Oct. 15, 1904 
480 Oct. 22, 1904 
504 Oct. 29, 1904 
•Ottawa-Ontario Power Company, Western Elec- 
trician 27: 85 Aug. 11, 1900 

Port Huron Light and Power Company, Engineering 

Record 49: 458 April 9, 1904 

The Puyallup River Water Power Development, 

Engineering News 52: 273 Sept. 29, 1904 

Engineering Record 50: 399 Oct. 1, 1904 

Saut Mortier Transmission Plant, American Elec- 
trician 17: 457 Sept., 1905 

The Sill Hydraulic Power Plant near Innsbruck, 

Engineering Record 52: 13 July 1, 1905 

Standard and Bay Company Plant, Engineer 

(United States) 40: 46 Jan. 1, 1903 

Snoqualmie Falls, American Electrician 13: 497 Oct., 1901 

Power 19: 1 Oct., 1899 

. . . . . 44: 398 Dec. 13, 1900 



802 STEAM POWER PLANT ENGINEERING 

Hydraulic — Continued. 

Swiss Combined Water Power and Gas Engine 

Plant, American Electrician 15: 113 

St. Croix Power Plant, Engineering Magazine 20: 954 

Sault Sainte Marie Power Plant (H. Von Schon), 

Engineering Magazine 24: 273 

Combined Hydraulic and Steam Plant, Stuyvesant 

Falls, New York, Engineering Record 43 : 3 

Sterling Hydraulic Company, Engineering Record . . 52 : 688 
St. Maurice-Lausanne, Switzerland, Western Elec- 
trician 33: 477 

Sewanee Falls Power Plant, Engineering Record ... 53 : 44 

Shawinigan Falls, Canada, Cassier's 26: 202 

Engineering Record 41 : 391 

Hydro-Electric Development at Turner's Falls, 

Electrical World 46: 263 

Steam Engine Plants. 

Aurora, Elgin and Chicago Railway, Street Railway 

Journal 19 : 143 

20:574 
Augusta Railway Power Station, Street Railway 

Journal 21 : 24 

Brooklyn Rapid Transit Company, Electrical 

World 46: 519 

Engineering Record 52 : 254 

Street Railway Journal 21 : 256 

26:432 

Street Railway Review 14 : 332 

Berlin Elevated and underground Railway, 

Engineering Record 45 : 389 

Boston Elevated Power Station, Street Railway 

Journal 17: 253 

20: 118 
Boston and Manhattan Railway, Street Railway 

Journal 20: 554 

Some British Central Electric Power Stations 

(H. F. Parshall), Cassier's 26: 233 

Breslau, Germany, Power 24: 742 

Brigham City, Utah, Light Plant, Engineering 

News 44: 235 

Boston Elevated, Lincoln Wharf Station, Engineer 

(United States) 40: 4 

Berlin Electric Railway, Engineer (United States) . . 40: 16 
Brussels Electric Power Station at Anderlecht, 

Engineer (United States) 41 : 555 

Boston and Worcester Street Railway, Engineer 

(United States) 40: 507 

Engineering Record 52 : 438 



March, 1903 
Feb., 1901 

Nov., 1902 

Jan. 5, 1901 
Dec. 16, 1905 

Dec. 20, 1903 
Jan. 13, 1906 
July, 1904 
April 29, 1900 

Aug. 12, 1905 



Feb., 1902 
Oct. 4, 1902 

Jan. 3, 1903 

Sept. 23, 1905 
Sept. 23, 1905 
Feb. 14, 1903 
Sept. 23, 1905 
May 20, 1904 

April 26, 1902 

March, 1901 
July 26, 1902 

Oct. 4, 1902 

June, 1904 
Dec, 1904 

Sept. 7, 1905 

Jan. 1, 1903 
Jan. 1, 1903 

Aug. 15, 1904 
Oct. 14, 1905 

July 1, 1903 



APPENDIX A 



803 



Steam Engine Plants — Continued. 

Brooklyn Navy Yard, Power 

Berkshire Street Railway Company, Street Railway 

Review 

American Electrician 

Engineering Record 

Brussels Electric Tramways, Power 

Bloomington and Normal Railway, Street Railway 

Journal 

Cranford, New Jersey, Engineering Record 

Consolidated Traction Company, Pittsburg, 

Street Railway Journal 

Street Railway Review 

Cleveland Electric Railway Company, Street 
Railway Journal 



Conneaut and Erie, Street Railway Journal 

Canton and Akron, Street Railway Journal 

Cleveland, Painesville, and Ashtabula, Street Rail- 
way Journal 

Chicago and Western Indiana, Western Elec- 
trician 

Cincinnati, Georgetown and Portsmouth Railway, 
Engineer (United States) 

Clyde Valley Electric Power Company, Engineer 
(United States) 

Capitol Traction Company, Washington, District 

of Columbia, Power 

Engineering Record 

Street Railway Journal 

Citizens' Light and Power Company, Power 

Chicago Electric Traction Company, Street Rail- 
way Review 

Engineering News 

Chattanooga Electric Company, Street Railway 
Review 

Carville (England) Power Plant, Street Railway 
Journal 

Dublin United Tramways, Power 

Detroit, Ypsilanti and Ann Arbor Railway, Street 
Railway Review 

Des Moines, Iowa, Engineering Record 

Street Railway Journal 

Denver Tramway System, Street Railway Journal . . 

Dayton and Muncie Electric Railway, Street 
Railway Journal 

Detroit United Railway, Street Railway Journal . . . 



20 


:21 


Dec, 1900 


12:813 


Nov. 20, 1902 


15:387 


Aug., 1903 


46:74 


July 26, 1902 


25: 1 


Jan., 1905 


25:934 


May 27, 1905 


49:355 


March 19, 1904 


15: 127 


March, 1899 


9:135 


Feb. 15, 1899 


15: 199 


April, 1899 


267 


May, 1899 


17:655 


June, 1901 


19:500 


April, 1902 


23:162 


Jan. 30, 1904 


23: 195 


Feb. 6, 1904 


23:805 


May 28, 1904 


24:93 


July 16, 1904 


37:425 


Dec. 2, 1905 


40: 353 


May 15, 1903 


42:524 


Aug. 1, 1905 


19 


1 


Feb., 1899 


39 


99 


Dec. 31, 1898 


15 


9 


Jan., 1899 


22 


1 


Feb., 1902 


8:184 


March 15, 1898 


39:2 


Jan. 6, 1898 


15: 136 


March 15, 1905 


21:902 


June 20, 1903 


19:1 


Jan., 1899 


10:5 


Jan. 15, 1900 


46:25 


July 12, 1902 


22:54 


July 11, 1903 


22:683 


Oct. 10, 1903 


26:979 


Dec. 2, 1905 


20 


•445 


Oct. 4, 1902 



804 



STEAM POWER PLANT ENGINEERING 



Steam Engine Plants — Continued. 

Everett Railway and Electric Company, Engineer 

(United States) 40: 839 Nov. 16, 1903 

Exeter, Hampton and Amesbury Street Railway, 

Street Railway Review 10: 630 Nov. 15, 1900 

Edison Electric Company, Atlantic Avenue 
Station, Boston, Transactions American Society of 

Mechanical Engineers 23 : 569 May, 1902 

Glasgow Municipal Tramways, Power 22: 1 Aug., 1902 

Street Railway Journal 15: 247 April, 1899 

17: 625 June, 1901 

Georgetown, Colorado, Western Electrician 27: 282 Nov. 3, 1900 

Grand Rapids, Holland and Lake Michigan Rail- 
way, Engineer (United States) 39: 179 March 15, 1902 

Hanover (Germany) Power Station, Power 25: 583 Oct., 1905 

Western Electrician 38: 55 Jan. 20, 1906 

Hartford and Springfield Railway, Engineering 

Record 45: 485 May 24, 1902 

Halifax, England, Tramway, Street Railway 

Journal 20: 293 Sept. 6, 1902 

Harrisburg, Pennsylvania, Street Railway Journal. 27: 61 Jan. 6, 1906 

Hazleton, Pennsylvania, Street Railway Journal ... 21 : 350 March 7, 1903 

Hebron, Ohio, Street Railway Journal 22: 147 Aug. 1, 1903 

Interborough Rapid Transit Company, American 

Electrician 16: 501 Oct., 1904 

Engineering Record 51 : 345 March 25, 1905 

50:384 Oct. 1, 1904 

424 Oct. 8, 1904 

456 Oct. 15, 1904 

490 Oct. 22, 1904 

510 Oct. 29, 1904 

541 Nov. 5, 1904 

Western Electrician 35:211 Sept. 17, 1904 

Power 24: 511 Sept., 1904 

Indianapolis and Cincinnati Traction Company, 

Engineering Record 51 : 329 March 18, 1905 

Street Railway Journal ... . 25: 502 March 18, 1905 

Interurban Railway and Terminal Company, 

Cincinnati, Engineer (United States) 41 : 251 April 1, 1904 

Independent Electric Light and Power Company, 

San Francisco, Power 22: 1 March, 1902 

Johnstown, Pennsylvania, Power Station, Street 

Railway Review 11: 426 July 15, 1901 

Johnston and Gloversville Railway, Street Railway 

Review 22: 308 Aug. 22, 1903 

Kankakee Power and Light Plant, Western Elec- 

#■-. trician 26:127 March 3, 1900 

Kingsbridge Power Station, Engineering Record. . . 50: 10 July 2, 1904 

Engineering News 43: 189 March 22, 1900 

Central Station at Kien, Engineer (United States) . . 42: 337 May 15, 1905 * 



APPENDIX A 805 

Steam Engine Plants — Continued. 

Kaw River Power Station, Kansas City, Street 

Railway Review 10: 560 Oct. 15, 1900 

Louisville and Nashville Railway Company, South 

Louisville, Kentucky, Engineer (United States) . . 42: 499 Aug. 1, 1903 
Louisville Railway Company, Engineer (United 

States) 42:511 Aug. 1, 1905 

Street Railway Review 9 : 439 July 15, 1899 

LaBella Power Plant, Goldfield, Colorado, Western 

Electrician 25: 267 Nov. 4, 1899 

Lisbon Railway Power Station, Street Railway 

Journal 17: 293 March, 1901 

Liverpool Electric Railway, Street Railway 

Journal 17: 390 April, 1901 

London Power Stations, Street Railway Journal .... 18 : 287 Sept., ' 1901 
Lackawanna and Wyoming Valley Railroad, 

Engineering Record 47: 624 June 13, 1903 

Street Railway Journal 21 : 868 June 13, 1903 

London County Council Tramway, Street Railway 

Journal 21: 821 June 6, 1903 

Leicester Power Station, Street Railway Journal .... 23 : 832 June 4, 1904 
Manchester Traction Company (W. V. Batson), 

Engineering Record 47: 107 Jan. 24, 1903 

Street Railway Journal 20 : 300 Sept. 6, 1902 

25:817 May 6, 1905 

Muskegon, Michigan, Light and Traction Company, 

Engineering Record 48: 452 Oct. 17, 1903 

Metropolitan West Side Elevated, Chicago, Street 

Railway Journal 15: 581 Sept., 1899 

709 
MeKeesport and Connellsville Railway, Engineering 

Record 48:264 Sept. 5, 1903 

Manx Railway, Isle of Man, Street Railway 

Journal 23: 357 March 5, 1904 

Milwaukee Gas Light Company, Engineering News. 55: 28 Jan. 11, 1906 
Moulineaux Power Station, Paris, American Elec- 
trician 16: 325 July, 1904 

Power Station of the Metropolitan Railway, 

Engineer (United States) 37: 271 Nov. 15, 1900 

American Electrician 12 : 112 March, 1900 

Monterey Gas and Electric Company, Engineer 

(United States) 42:87 Jan. 16, 1905 

Market Street Railway Company of San Francisco, 

Power 19:3 March, 1899 

Manhattan Elevated Railway, Power 21: 1 April, 1901 

Street Railway Review 9: 82 Feb. 16, 1899 

Engineering (United States) 39:80 Feb. 1, 1902 

Engineering News 47: 82 Jan. 30, 1902 

Milford, Attleboro and Woonsocket Railway, 

Street Railway Review 10: 638 



806 



STEAM POWER PLANT ENGINEERING 



Steam Engine Plants — Continued. 

New York Central Power House, Electrical World ... 46 : 95 

Power 24: 228 

Newcastle and District Electric Lighting Company, 

Engineer (United States) 40: 29 

Northwestern Elevated Railroad, Street Railway 

Review 15 : 199 

267 
17: 655 
19:500 
23: 162 

Newcastle, England, Engineering Record 46: 157 

New York Gas and Electric Light and Power 

Company, Engineering News 45: 375 

Oakland, California, Engineering Record 49: 591 

Omaha and Council Bluffs Railway and Bridge 

Company, Street Railway Review 9 : 293 

Paris Metropolitan, American Electrician 16: 111 

Engineer (United States) 40 : 13 

Pan-American " Exposition, Engineer (United 

States) 38: 158 

Paris Exposition Plant, Power 19: 7 

Pittsburg Traction Company, Street Railway 

Review 9 : 135 

15: 401 
Philadelphia Rapid Transit Company, Street 

Railway Review 15: 526 

Engineering Record 47: 611 

52:340 
Electrical Power Development at Portland, 

Oregon, Electrical World 46 : 174 

Engineering Record 52 : 142 

Peekskill (New York) Light and Railway Company, 

Street Railway Journal 20 : 92 

Pacific Electric Railway Company, Los Angeles, 

Street Railway Journal 23 : 394 

Richmond, Virginia, Engineering Record 49: 11 

Rock Island Railway Shops at East Moline, 

Illinois, Engineering Record 50: 137 

Rochester Power Plant, Engineer (United States). 38: 312 
Scioto Valley Traction Company, Engineering 

Record 50: 644 

St. Louis Exposition Power Plant, Engineering 

News 42: 320 

American Electrician 16: 528 

Engineering News 42 : 223 

Western Electrician 35 : 303 

Power 23: 481 

Saginaw and Bay City Light Company, Engineer 

(United States) 42: 399 



July, 1905 
April, 1904 

Jan. 1, 1903 



Aug. 16, 1902 

May 23, 1901 
May 7, 1904 

May 15, 1899 
March, 1904 
Jan. 1, 1903 

June 1, 1901 
Dec, 1899 

Feb. 15, 1899 
July 15, 1905 

Sept. 15, 1905 
June 20, 1903 
Sept. 23, 1905 

July 29, 1905 
Aug. 5, 1905 

July 19, 1902 

March 12, 1904 
Jan. 2, 1904 

July 30, 1904 
Sept. 1, 1901 

Dec. 3, 1904 



Oct., 1904 
Sept. 15, 1904 
Oct. 15, 1904 
Nov., 1903 

June 15, 1905 



APPENDIX A 



807 



rEAM Engine Plants — Continued. 

St. Louis Transit Company, Street Railway 

Journal 11: 813 Nov. 15, 1901 

Engineering News 47: 269 April 3, 1902 

297 April 10, 1902 

Stark Electric Railway Company, Street Railway 

Journal 25: 10 Jan. 7, 1905 

The Sunderland District Tramways, Street Railway 

Journal 26: 96 July 15, 1905 

St. Clair Tunnel, Grand Trunk Railway, Engineer- 
ing News 55: 62 Jan. 18, 1906 

Dayton, Springfield and Urbana Railway, American 

Electrician 12: 415 Sept., 1900 

Southern Pacific Company, Galveston, Texas, 

American Electrician 15: 476 Oct., 1903 

Syracuse Power Station, Street Railway Journal . ... 19: 517 May, 1902 

Sydney, New South Wales, Street Railway Journal . . 20 : 930 Dec. 6, 1902 

Seattle Power Station, Street Railway Journal 21 : 649 May 2, 1903 

South Side Elevated, Western Electrician 26: 231 April 14, 1900 

Springfield, Illinois, Light and Power Company, 

Electrical World 47: 253 Feb. 3, 1906 

Toledo and Monroe, Street Railway Journal 18: 124 Aug., 1901 

Tokio Electric Railway, Street Railway Journal ... 19 : 245 March, 1902 

Toledo, Port Clinton and Lakeside Railway, 

Engineer (United States) 43: 295 March, 1906 

Toledo and Western Railway, Street Railway 

Journal 20: 980 Dec. 20, 1902 

Twin City Rapid Transit Company, Engineering 

Record 50: 692 Dec. 10, 1904 

Street Railway Review 14 : 441 July 20, 1904 

Terminal Railroad Association of St. Louis, 

Engineering Record 51 : 92 Jan. 28, 1905 

Trans-Mississippi Exposition, Omaha, Power 18: 7 June, 1898 

Toledo and Maumee Valley Railway, Street Rail- 
way Review 9: 613 Sept. 15, 1899 

American Electrician 12: 464 Oct., 1900 

United Railways and Electric Company, Balti- 
more, Engineer (United States) 39: 357 May 15, 1902 

Union Traction Company, Street Railway Review . . . 9: 12 Jan. 15, 1899 

Union Traction Company, Anderson, Indiana, 

Engineering Record 43: 495 May 25, 1901 

Street Railway Journal 18: 826 Dec, 1901 

United Electric Company, Hoboken, New Jersey, 

Engineering Record 46: 56 July 19, 1902 

Union Loop Power House, Chicago, Western Elec- 
trician 24: 1 Jan. 7, 1899 

Virginia Electric Railway and Equipment Com- 
pany, Engineering News 45: 318 May 2, 1901 

Voltellina Three-Phase Railway, Street Railway 
Journal 21 : 791 May 30, 1903 



308 



STEAM POWER PLANT ENGINEERING 



Steam Engine Plants — Continued. 

Worcester and Blackstone Valley Street Railway, 

Engineering Record 47: 462 May 2, 1903 

Warren and Jamestown Single-Phase Railway, 

Street Railway Journal 27: 270 Feb. 16, 1906 

Electrical World 47: 363 Feb. 17, 1906 

Washtenaw Electric Company, Engineer (United 

States) ,. . . 38: 150 May 15, 1901 

Willimantic Gas and Electric Company, Engineer 

(United States) 42: 163 March 1, 1905 

West Albany Power Station, Power 24: 677 Nov., 1904 

Wilkesbarre Interurban Railway, Street Railway 

Review 9:131 Feb. 15, 1899 

Weehauken Station New York Central and 

Hudson River Railroad, American Electrician. . 17: 501 Oct., 1905 

Engineering News 54: 506 Nov. 16, 1905 

Street Railway Journal 26: 872 Nov. 11, 1905 

Engineering Record 52: 553 Nov. 11, 1905 

Western Pennsylvania Railway and Light System, 

Street Railway Journal 20: 139 Aug. 2, 1902 

Zanesville Railway and Light Company, Electrical 

World 43: 500 March 19, 1904 

Turbine Plants. 

Amsterdam-Haarlem Electric Railway, Engineer- 
ing Record 51 : 18 Jan. 17, 1905 

Street Railway Journal 25: 21 Jan. 7, 1905 

Brooklyn Rapid Transit, Williamsburg Plant, 

Street Railway Journal 26: 432 Sept. 23, 1905 

Generating Station Champ, American Electrician. . 17: 65 Feb., 1905 
Clyde Valley Electric Power Company, Engineering 

Record 52: 209 Sept. 9, 1905 

Chelsea Generation Station of the London Under- 
ground, Engineering Record 52: 215 Feb. 25, 1905 

Power 23: 421 Aug., 1903 

Street Railway Journal 25: 388 March 4, 1905 

Commonwealth Electric Company, Fisk Street 

Station, Western Electrician 38: 55 Jan. 20, 1906 

Power 26:715 Dec, 1906 

Dubuque, Iowa, Power Plant, Engineering Record. 20: 202 Aug. 13, 1904 

Detroit Edison Company, Engineering Record 52 : 194 Oct. 7, 1905 

DeBeers Consolidated Mines, Engineering Record .. . 51:4 . Jan. 7, 1905 
Edison Electric Illuminating Company, Boston, 

Engineering Record 51: 150 Feb. 11, 1905 

Power 25: 389 July, 1905 

Hartford Electric Light Company, Dutch Point 

Plant, Engineering Record 51 : 204 Feb. 25, 1905 

Long Island Railroad Power House, Street Railway 

Journal 25: 24 Jan. 7, 1905 

Engineering Record 49: 454 April 9, 1904 



APPENDIX A 809 

Turbine Plants — Continued. 

Los Angeles Edison Company, Power 26: 67 Feb., 1906 

Mexican Central Shops at Aguascalientes, Engineer- 
ing Record 50: 227 Aug. 20, 1904 

247 Aug. 27, 1904 
Municipal Turbine Plant, Anderson, Indiana, 

Engineer (United States) 42: 641 Oct. 2, 1905 

Manchester, England, Power Station, Street Rail- 
way Journal 25: 934 May 27, 1905 

New York Edison Waterside Station, Electrical 

World 46:383 Sept. 2, 1905 

435 Sept. 9, 1905 

Power 22:1 Jan., 1902 

New York Central Steam-Electric Station, Power 26: 131 March, 1906 

Engineer (United States) 43: 733 Nov. 15, 1906 

New York and Long Island River Power Station, 

Power 26: 199 April, 1906 

New Orleans Power House, Power 24: 651 Nov., 1904 

Street Railway Review 9: 393 June 15, 1899 

New Bedford Power Station, Street Railway 

Review 11: 884 Dec, 1901 

Old Colony Street Railway, Quincy Point Station, 

Engineering Record 51 : 646 June 10, 1905 

Street Railway Journal 25: 1022 June 10, 1905 

Engineer (United States) 43; 85 Jan. 15, 1906 

Steam Turbine Plant at Poughkeepsie, New York, 

Shop, Engineering Record 51 : 454 April, 1905 

Potomac Electric Power Company, Washington, 

District of Columbia, Power 27: 277 May, 1907 



Isolated Stations. 

Apartments. 

Collingwood Apartment Hotel, New York, 

Engineering Record 45: 323 April 5, 1902 

Ansonia Apartment, New York, Engineering 

Record 46: 467 Nov. 15, 1902 

Hospitals. 

Agnes Memorial Sanatorium, Engineering Record ... 50: 312 Sept. 10, 1904 

Connecticut Hospital for the Insane, Engineering 

Record 52: 44 July 8, 1905 

Lakeside Hospital, Cleveland, Ohio, Engineer 

(United States) 39 : 108 Feb. 15, 1902 

Massachusetts General Hospital, Boston, Transac- 
tions American Society of Mechanical Engineers 22:392 Jan., 1901 

Hotels. 

Bellevue-Stratford Hotel, Philadelphia, Engineer- 
ing Record 51:14 Jan. 7, 1905 



810 STEAM POWER PLANT ENGINEERING 

Hotels — Continued. 

Hotel Belmont, New York, Engineering Record 52: 739 Dec. 30, 1905 

53:9 Jan. 6, 1906 

56 Jan. 13, 1906 

81 Jan. 20, 1906 

Hotel Gotham, New York, American Electrician 17: 551 Nov., 1905 

Engineering Record 52: 517 Nov. 4, 1905 

New York Athletic Club, Electrical World 31: 463 April 16, 1898 

585 May 14, 1898 

University Club, Engineering Record 46: 36 July 12, 1902 

Manufacturing Plants. 

American Steam Pump Company, Engineer 

(United States) 39: 217 April 1, 1902 

Atlas Knitting Mills, Power 25: 408 July, 1905 

American Lithographic Company, American 

Electrician 10 : 1 

Armour Packing Company (J. E. Smith), American 

Electrician 12: 202 May, 1900 

Western Electrician 23: 305 Nov. 26, 1898 

Anheuser-Busch Brewing Association, Engineer 

(United States) 42: 67 Feb. 1, 1905 

Power 25: 84 Feb., 1905 

Booth Cold Storage, Engineering Record 45: 97 Feb. 1, 1902 

Columbian Cordage Company, Engineering Record. 50: 447 Oct. 15, 1904 

Power Plants for Cotton Mills, Engineer (United 

States) 41: 135 Feb. 15, 1904 

Cascade Water Power and Light Company, Cas- 
cade, Canada, Engineer (United States) 40: 603 Aug. 1, 1903 

Camden Interstate Railway Company, Street 

Railway Review 15: 269 May 15, 1905 

DeCew Falls Power Plant, Western Electrician 38: 115 Feb. 10, 1906 

DeSabla, California, Station, Engineering News . ... 54: 131 Aug. 10, 1905 

Dan River Power and Manufacturing Company's 

Plant, Engineering Record 50: 291 Sept. 3, 1904 

Power Plants of Edison Electric Company of Los 

Angeles, Engineering Record 51 : 211 Feb. 25, 1905 

302 March 11, 1905 

325 March 18, 1905 

Electric Generating Station at Glomen, near 

Kykkelsrud, American Electrician 17: 409 Aug., 1905 

Engineering Record 50: 19 July 2, 1904 

Elgin Watch Works, Elgin, Illinois, Engineering 

Record 56: 294 Sept. 14, 1907 

Hydro-Electric Power Development for Guana- 
juato, Mexico, Engineering Record 50: 195 Aug. 13, 1904 

General Electric Company's Hudson River Plant, 

Electrical World 43: 1115 June 11, 1904 

Garvins Falls, New Hampshire, Engineering 

Record 49: 668 May 28, 1904 



APPENDIX A 811 

Manufacturing Plants — Continued. 

Gresivaudan Valley Power Plant, Western Elec- 
trician 

Hudson River Company, Mechanicsville, New York 23: 168 Oct. 1, 1898 

Engineering News 40: 130 Sept. 1, 1898 

Western Electrician 23: 135 Sept. 3, 1898 

Hampton, Virginia, Engineering Record 48: 179 Aug. 15, 1903 

Indian Power Plant, Engineer (London) 101 : 36 Jan. 12, 1906 

Engineering 81 : 103 Jan. 19, 1906 

Jhelum Power Plant, Engineer (United States) 43 : 240 March 15, 1906 

Kern River Company's Enterprise, Engineering 

News 52:55 July 21, 1904 

Little Falls, Montana, Power Station, Engineering 

Record 51:616 June 3, 1905 

McCormick Twine Mills, Western Electrician 28: 109 Feb. 16, 1901 

National Cash Register Company, Dayton, 

Engineer (United States) 39 : 136 March 1, 1902 

New England Confectionery Company, Engineer 

(United States) 41:231 April 1, 1904 

Otis Elevator Company, Yonkers, New York, 

Engineer (United States) 39: 327 May 1, 1902 

Olympia Cotton Mills, Engineer (United States) ... 38: 382 Oct. 15, 1901 

Power Plant of a Large Silk Mill, Engineer (United 

States) 38: 424 Nov. 15, 1901 

Power Plant of a Sulphite Mill, Engineer (United 

States) 38: 445 Dec. 1, 1901 

Swift and Company, American Electrician 13: 149 April, 1901 

Stickney and Poor Spice Factory, Charlestown, 

Massachusetts, American Electrician 14: 207 May, 1902 

United Machinery Company, Engineering Record. . 52: 198 Aug. 15, 1905 

Whittall Mills, Engineering Record 51 : 510 May 6, 1905 

Washington Mill Company, Lawrence, Massa- 
chusetts, Engineering 66: 533 Oct. 21, 1898 

Western Electric Company, New York Factory, 

American Electrician 12: 333 July, 1900 

Wetmore Tobacco Factory, American Elec- 
trician 13: 61 Feb., 1901 

Western Wheel Works, Chicago, Western Elec- 
trician 22: 75 Feb. 5, 1898 

Wood Worsted Mill, Lawrence, Massachusetts, 

Power 27: 73 Feb., 1907 

Miscellaneous. 

Boston South Terminal Station, Engineering 

Record 39: 346 March 18, 1899 

Brooklyn Institute of Arts and Science, American 

Electrician 17: 121 March, 1905 

Birmingham University (C. A. Smith), Engineering . 80: 341 Sept. 15, 1905 

397 Sept. 29, 1905 

507 Oct. 20, 1905 



812 



STEAM POWER PLANT ENGINEERING 



Miscellaneous — Continued. 

Bryn Mawr College (C. G. Gray), Engineering 

Record 53: 183 Feb. 17, 1906 

University of California, Engineer (United States) . . 42: 263 April 15, 1905 

Chicago, Milwaukee and St. Paul Railway Shops, 

Engineering Record 48: 594 Nov. 14, 1903 

Columbia University, Engineering Record 39: 546 May 13, 1899 

University of Chicago, Engineering Record 45: 246 March 15, 1902 

DuBois Shops, Engineering Record 46: 218 Sept. 6, 1902 

Street Railway Journal 21 : 694 May 9, 1903 

Elizabethport Railroad Shops, Engineering Record . . 45: 581 June 21, 1902 

Hippodrome, New York, Engineering Record 52: 229 Aug. 26, 1905 

Harvard Electric Light Plant (W. L. Robb), 

American Electrician 12 : 107 March, 1900 

Lackawanna Railroad, Engineering Record 52: 507 Nov. 4, 1905 

Michigan University (H. S. Carhartv), Electrical 

World 31 : 550 May 7, 1898 

Nassau, Bahama Islands, Plant, Electrical 

World 37: 914 June 1, 1901 

Ohio State University, Engineer (United States) . . 39: 164 March 15, 1902 

Electrical World 34: 1005 Dec. 30, 1899 

Princeton University, Engineer (United States) ... 41: 411 June 15, 1904 

Simmons College, Boston, Engineering Record 51: 161 Feb. 11, 1905 

Scranton Schools, Engineer (United States) 37: 148 June 1, 1900 

Power Plant of a University (E. A. Darling), 
Transactions American Society of Mechanical 

Engineers 20: 663 1899 

Piping Plans for the Onondaga County Court- 
House, Syracuse, New York, Power 26 : 1 Jan., 1906 

United States Bureau of Engraving, Iron Age 79 : 34 Jan. 3, 1907 

Office Buildings. 

Arthur Building, New York, Engineering Record. . 50: 725 Dec. 17, 1904 

Atlantic Building, New York, Engineer (United 

States) 38: 262 July 1, 1901 

Arcade Building, Dayton, Ohio, Engineering 

Record 49: 767 June 18, 1904 

Broadway Exchange Building, New York, Engineer 

(United States) 38: 277 Aug. 1, 1901 

Engineering Record 45: 374 Aug. 1, 1901 

Columbia Office Building, Engineering Record 48 : 103 

Cable Building, Chicago, Western Electrician 25: 377 

Chicago and Northwestern Railway Office Building, 

Engineer (United States) 42: 739 Nov. 15, 1905 

Commercial National Bank Building, Chicago, 

Engineer (United States) 44 Dec. 2, 1907 

Commerce Realty Building, St. Louis, Engineering 

Record 39: 33 Dec. 10, 1898 

Ellicott Square Building, St. Louis, Electrical 

World 31: 519 



APPENDIX A 813 

Office Buildings — Continued. 

Equipment of Tall Office Buildings, in New York, 

Engineering Record 39: 550 

Frick Building, Pittsburg, Engineering Record 45: 459 

Farmers' Bank Building, Pittsburg, Engineering 

Record 47: 492 

First National Bank Building, Chicago, Engineer- 
ing Record 54:312 Sept. 22, 1906 

360 Sept. 29, 1906 

Flat-iron Building, New York, Engineer (United 

States) 40: 296 April, 1903 

Federal Building, San Francisco, Engineering 

Record 47: 407, 578 April 18, 1903 

Heyworth Building, Engineer (United States) 42: 611 Sept. 15, 1905 

Isolated Plant for Office Buildings, Electrical 

World 32: 108 

Isolated Light and Power Plant, Western Elec- 
trician 24: 910 

Consolidated California and Virginia Mining Com- 
pany, American Electrician 14: 425 Sept., 1902 

Keystone Bank Building, Pittsburg, American 

Electrician 15: 171 

Kimball Building, Boston (H. S. Knowlton), 

Engineer (United States) 42: 819 Dec. 15, 1905 

Large and Modern Isolated Plant, American 

Electrician 15:1 

Land Title and Trust Building, Philadelphia, 

Electrical World ■ 32 : 45 

Murphy Power Building, Engineer (United States). 42: 67 Jan. 16, 1905 

Mutual Life Building, Engineering Record 47 : 85 

Maiden Lane Building, New York, Engineering 

Record 48: 770 

Methodist Book Concern, Chicago, Western Elec- 
trician 22: 199 

Metropolitan Life Building, New York, Engineer- 
ing Record 55: 97 Jan. 26, 1907 

Modern Commercial Building, Electrical World 32 : 623 

National Bank Building, Pittsburg, Engineer 

(United States) 40: 387 June 1, 1903 

First National Bank, Chicago, Power 25: 297 May, 1905 

First National Bank, Uniontown, Pennsylvania, 

Engineering Record 46: 13 

New Building, Dallas, American Electrician 14:11 Jan., 1902 

Oliver Building, Boston, Engineering Record 50: 717 Dec. 17, 1904 

Power Plants of Office Buildings, Engineering 

News 51: 537 

Engineering Record 49 : 725 

Power 24: 419 July, 1904 

Transactions American Society of Mechanical 

Engineers 20: 880 July, 1899 



814 



STEAM POWER PLANT ENGINEERING 



Office Buildings — Continued. 

Power Plants of the Tall Office Buildings (J. H. 

Wells), Engineering 78: 130 July 22, 1904 

Engineering Review 14: 19 July, 1904 

Power Plants of Tall Office Buildings (Wells and 
Bolton), Transactions American Society of Mechan- 
ical Engineers 25: 685 June, 1904 

Phipps Building, Pittsburg, Engineering Record . ... 50: 343 Sept. 17, 1904 

Power Building, in Providence, Engineering Record 51: 162 Feb. 11, 1905 

Park Row Building, Power 22: 1 Oct., 1902 

Electrical World 34: 5 

Prudential Building, Newark, New Jersey, 

Engineering Record 46 : 367 

Pittsburg and Lake Erie Terminal, Engineering 

Record 46: 98 Aug. 2, 1902 

152 Aug. 16, 1902 

Pennsylvania Railroad Station, Pittsburg, 

Engineering Record 46: 203 Aug. 30, 1902 

Rose Building, Cleveland, Engineer (United States) 38: 404 Nov. 1, 1901 

Railway Exchange Building, Chicago, Engineer 

(United States) 41 : 763 Nov. 15, 1904 

Rogers-Peet Building, New York, Engineering 

Record 46: 36 July 12, 1902 

Republic Building, St. Louis, American Elec- 
trician 12: 67 Feb., 1900 

Rock Island Station, Chicago, Engineering Record . . 48:328 Sept. 19, 1903 

Tribune Building, Chicago, Engineering Record .... 45 : 607 

Western Reserve Building, Engineer (United 

States) 38: 87 March 1, 1901 

Wells Building, Milwaukee, Engineer (United 

States) 40: 189 March 2, 1903 

Stores. 

Boston Post Office, Electrical World 46: 486 Sept. 16, 1905 

Boston Store, Chicago, Engineer (United States) ... 44: 559 June 15, 1907 

Daniels and Fisher, Denver, Engineering Record . . . 50: 294 Sept. 3, 1904 

Ferguson-McKinney Dry Goods Company, St. 

Louis, Engineer (United States) 40: 131 Feb. 2, 1903 

Marshall Field and Company, Engineering Record ... 48: 366 Sept. 26, 1903 

Western Electrician 31 : 165 Sept. 13, 1902 

New York Federal Building, Electrical World 31 : 379 March 25, 1896 

Government Printing Office, Electrical World 31: 94 Jan. 15, 1898 

115 Jan. 22, 1898 

Engineering Record 47: 512 May 16, 1903 

543 May 23, 1903 

Modern Printing Plant, Engineer (United States) ... 41 : 443 July 1 , 1904 

Toetz Store, Munich, Engineering Record 53: 160 Feb. 10, 1906 

Wanamaker Store, Engineering Record 52: 94 July 22, 1905 

125 July 29, 1905 

53:219 ' Feb. 24, 1906 



APPENDIX A 



815 



Power Plant Design. 

Power Equipment for City Roads, Street Railway 

Journal 22: 358 

Production and Distribution of Alternating Current 

for Large City Systems, Street Railway Journal 22 : 506 

Discussion 483 

Oil Problem in Power Stations, Street Railway Journal 22 : 664 
Hydraulics in Connection with Street Railway 

Operation, Street Railway Journal 22 : 709 

Electric Power Generating and Transmission Station, 

Street Railway Journal 24 : 800 

Wire Glass for Power Stations, Street Railway Journal 24: 1049 
Stark Electric Railway Company, Street Railway 

Journal 25: 10 

Electrification of the London Underground Electric 
Railway Company's System (S. B. Fortenbaugh), 

Street Railway Journal 25 : 388 

Practical Operation and Maintenance of Electrical 

Equipment, Street Railway Journal 8: 33 

Importance of the Power House, Street Railway 

Review 9 : 458 

Transmission of Power, Street Railway Journal 11: 24 

11:154 
11:203 
11:363 
11:427 
11:497 
11:669 
11:831 
Communication in Electric Generating Stations, 

Street Railway Journal 11 : 910 

Power House Helps, Street Railway Journal 12 : 352 

Power Speed and Efficiency Curves, Street Railway 

Review 12: 534 

Power Transmission and Distribution in Utah, 

Street Railway Review 13: 288 

Power for Interurban Lines, Street Railway Review.. 13:808 
Power Plant Experiences, Street Railway Review . ... 14: 898 
Hints in Laying Out a Power Plant, American 

Electrician 13 : 495 

Power Plant of a Modern Telephone Exchange, 

American Electrician 17: 190 

The Power Station (F. N. Bushnell), American 

Electrician 17: 561 

The Design of a 1,500-Kilowatt Steam Electric 
Central Light and Power Plant (F. Koester), 

Electrical Review (New York) 50: 634 

670 
706 



Aug. 29, 1903 

Sept. 12, 1903 

Oct. 3, 1903 

Oct. 10, 1903 

Oct. 29, 1904 
Dec. 10, 1904 

Jan. 7, 1905 

March 4, 1905 

Jan., 1898 

July 15, 1899 
Jan. 15, 1901 
March 15, 1901 
April 15, 1901 
June 15, 1901 
July 15, 1901 
Aug. 15, 1901 
Oct. 10, 1901 
Nov. 10, 1901 

Dec. 15, 1901 
June 20, 1902 

Sept. 20, 1902 

May 20, 1903 
Oct. 20, 1903 
Nov. 20, 1904 

Oct., 1901 

April, 1905 

Nov., 1905 



April 20, 1907 
April 27, 1907 
May 4, 1907 



816 STEAM POWER PLANT ENGINEERING 

Suggestions for Improvement in Power Plants 

(A. Bement), Cassier's 13: 538 April, 1898 

Tendencies in Power Plant Design (G. L. Clark), 

Cassier's 29 : 433 March, 1906 

Standardizing the Power Plant, Electrical Age 32 : 317 May, 1904 

Unipolar Dynamos and Modern Central Station 

Design, Electrical Age 34: 114 Feb., 1905 

Making the Most of an Old Plant, Electrical Age 34: 211 March, 1905 

Duplication of Electrical Apparatus in Power 

Houses, Electrical Age 34: 443 June, 1905 

Design of Power Stations, Electrical World 43: 991 May 21, 1904 

1030 May 28, 1904 

1090 June 4, 1904 

Design of an Isolated Power and Lighting Plant, 

Electrical World 47: 372 Feb. 17, 1906 

Piping a Steam Plant, Engineer (United States) 36: 78 April 1, 1899 

117 May 15, 1899 

Power Plant Specifications, Engineer (United 

States) 38: 293 Aug. 15, 1901 

Injudicious Design and Management of Power 

Plants, Engineer (United States) 40: 364 May 15, 1903 

Reversal of a Large Power Station, Engineer 

(United States) 40: 876 Dec. 1, 1903 

Power Plants for Cotton Mills, Engineer (United 

States) 41: 135 Feb., 1904 

Power Station Design, Engineering Magazine 19: 902 Sept., 1900 

Power Plant of Apartment Houses, Engineering 

Magazine 26: 713 Feb., 1904 

Power Station Design (Merz and McLellan), 

Engineering Magazine 27: 822 Aug., 1904 

Modern Power Plant Design (Franz Koester), 

Engineering Magazine 29: 689 Aug., 1905 

811 Sept., 1905 

30:71 Oct., 1905 

182 Nov., 1905 

Planning and Construction of the Power Plant 

(A. E. Dixon), Engineering Magazine (Serial) 31 : 722 Aug., 1906 

32:860 May, 1907 

Design and Construction of Modern Central Light 

Station Power House (H. H. Humphrey), 

Engineering News 43 : 35 Jan. 18, 1900 

Design of Piping for Electric Power House, 

Engineering Record 39 : 54 Dec. 17, 1898 

Power Plant in a Factory, Engineering Record 39 : 430 April 8, 1899 

Design and Construction of Central Station, 

Engineering Record 40: 651 Dec. 9, 1899 

Foundation of a Large Power Plant, Engineering 

Record 40: 681 Dec. 16, 1899 

Design of Steam Power Plants (H. C. Meyer), 

Engineering Record 41 : 597 June 23, 1900 



APPENDIX A 



817 



Design of Steam Power Plant Condensers (H. C. 

Meyer, Jr.), Engineering Record 43: 55 Jan. 19, 1901 

Feed Water Heaters and Economizers (H. C. Meyer, 

Jr.), Engineering Record 43: 224 March 9, 1901 

Mechanical Draft and Chimneys (H. C. Meyer, Jr.), 

Engineering Record 43: 468 May 18, 1901 

Some Departures in Design in a Detroit Central 

Power Station, Engineering Record 51: 356 March 25, 1905 

Cement in Central Station Design, Engineering Record 51:481 April 29, 1905 

Compactness in Power Plant Design, Engineering 

Record 51: 560 May 20, 1905 

Power Station Design (F. N. Bushnell), Engineering 

Record 52:460 Oct. 21, 1905 

Metamorphosis of an Electric Power Station, Power 18: 1 June, 1898 

Power Plant Location, Power 24: 646 Nov., 1,904 

24: 770 Dec, 1904 

Power Plant Location (H. D. Jackson), Power 25: 435 July, 1905 

Design of a 100-Horse-Power Steam Plant (J. F. 

Hobart), Power..., 27:421 July, 1907 

Present Trend of Station Practice, Street Railway 

Journal 19: 750 June 14, 1902 

Designing of Steam Power Plants (W. C. Kerr), Street 

Railway Journal 20: 497 Oct. 4, 1902 

Safety Devices in Central and Substations, Street 

Railway Journal 21 : 672 May 2, 1903 

Economical and Safe Limits on Size of Central 

Stations, Street Railway Journal 27: 671 May 2, 1903 

Design of City Power Stations, Street Railway Journal 22: 282 Aug. 29, 1903 

Evolution of Modern Power Station, Street Railway 

Journal 24: 555 Oct. 8, 1904 

Size of Power Stations, Street Railway Journal 26: 131 July 22, 1905 

Power Station (F. N. Bushnell), Street Railway Journal Sept. 30, 1905 

Underground Distribution of Power for Urban Electric 

Station (James Heywood), Street Railway Journal. . .26: 269 Aug. 19 1905 

Designing Boilers for a Small Street Railway Plant, 

Street Railway Review 9: 125 Feb., 1899 

9: 188 March, 1899 

9:263 April, 1899 

9: 190 1899 



Storage Batteries and Railway Power Stations, Street 
Railway Review 



Bibliography, 1908-1910. 

Gas Engine Plants. 

Bituminous Power Producer Plant: 

Pro. Engr. Soc. W. Penn 25: 603 Jan., 1910 

Engineering, London 108: 317 Sept. 24, 1909 

Engineering Record 59: 817 June 26, 1909 

Engineering News 61 : 42 Jan. 14, 1909 

Jour. A.S.M.E 31: 1343 Dec, 1909 

Blast Furnace Power Plant, Engineering and Mining 

Journal 87: 20 Jan. 2, 1909 



818 STEAM POWER PLANT ENGINEERING 

Gas Engine Plants— Continued. 

Boston Elevated R. R. Co.'s Plant, Engineering 

Record 59: 352 March 27, 1909 

Carnegie Steel Plant, Power and Engineer 28: 639 April 28, 1908 

Central Station at Harvard, 111., Electrical 

World 53: 913 April 15, 1909 

Central Station at Aberdeen, South Dakota, 

Electrical World 54 : 1567 Dec. 30, 1909 

Combined Steam and Gas Power Plant: Engi- 
neering Record 60: 329 Sept. 18, 1909 

Costs of a Gas Engine and of a Combined Steam 

Plant, Engineering Record 60: 272 Sept. 4, 1909 

Gas Engine in Central Station Work, Iron Age 84: 256 July 22, 1909 

Gas-Electric Plant at Gary, Indiana, Power and 

Engineer 29: 268 Aug. 18, 1908 

Western Electrician 43: 245 Oct. 3, 1908 

Gas-Driven Blowing Plant at Gary, Indiana, 

Engineering Record 59: 274 March 6, 1909 

Gas Power vs. Steam Power for Central Stations, 

Electrical Review 65: 764 Nov. 12, 1909 

Philadelphia Gear Works, American Machinery 32: 155 July 22, 1909 

Recent Development of the Producer Gas Power 

Plant in the United States, U. S. Geological 

Survey Bui. 416 No. 11142 

Small Isolated Producer Gas Plant, Power and 32: 68 Jan. 1, 1910 

Engineer 31: 697 Oct. 26, 1909 

Working Results, Gas-Electric Power Plant, 

Pro. Am. Inst. Elec. Engrs 27: 1123 July 1, 1909 

Hydro-Electric. 

Hennepin Power House, St. Anthony Falls, 

Engineering Record 59: 676 May 29, 1909 

La Crosse, Wis., Water Power Company, Elec- 
trical World 55: 801 March 31, 1910 

Lansing, Mich., Hydro-Electric with Steam 

Auxiliaries, Electrical Review, New York 55: 1235 Dec. 25, 1909 

Low-Head Hydro-Electric, Tippecanoe River, 

Monticello, Ind., Electrical World, 54: 975 Oct. 21, 1909 

Rainbow Falls, Engineering Record 61 : 292 March 12, 1910 

Sioux Falls, Electrical World 53: 963 April 22, 1909 

Telluride Power Company, Bear River Plant, 

Engineering Record 61 : 353 March 26, 1910 

Wenitchee Electric Company, Journal of Elec- 
tricity Power andGas 23: 87 July 31, 1909 

Steam Engine — Central Stations. 

Central Pennsylvania Traction Company, Elec- 
tric Railway Review 19: 455 April 11, 1908 

Commonwealth Power Company, Power and 

Engineer 31: 127 July 27, 1909 



APPENDIX A 

Steam Engine — Central Stations — Continued. 

Home-Electric and Steam Heating Company, 

Electrical World 54: 1345 

Indianapolis, Crawfordsville and Western Trac- 
tion System, Street Railway Journal 31 : 850 

Louisville Lighting Company, Power 30: 663 

Redonda Power Plant, Redonda, Cal., Electrical 
Railway Journal 32: 618 

United Railways of Baltimore, Street Railway 
Journal 31: 770 

Williamsport, Penn., New Plant at, Power and 

Engineering 31: 213 

Steam Turbine Plants. 

Brockton, New Lighting Station at, Power and 

Engineer 30: 315 

Commonwealth Edison Company, Chicago, 

Fisk Street, Electrical World 51 : 1023 

Northwest Station, Electrical World 55: 667 

Quarry Street, Electrical World 53: 17 

Double Flow Turbines at Brunat Island, Street 

Railway Journal 31 : 908 

Engineering Record 57 : 693 

Government Plant at Washington, Power and 

Engineer 32: 838 

Hartford Electric Light Company, Power and 

Engineer 31 : 884 

Hoboken Passenger Terminal of the Lacka- 
wanna Railway, Engineering Record 59: 184 

Jackson Electric Railway, Light and Power 

Company, Street Railway Journal 31 : 278 

New York Edison, Waterside No. 2, Electrical 

World 53: 545 

Knoxville Railway and Light Company, Electric 

Railway Journal 32 : 534 

Norfolk Traction Power Plant, Power and Engi- 
neer 32: 702 

Recent Electric Railway Power Station Design, 

Electrical Railway Journal 33: 538 

Sayre Electric Company, Electrical World 55: 273 

Isolated Stations. 

Apartment Buildings, Apthorp Apartment, New 

York, Engineering Record 60: 69 

Belnord Apartment, New York, Domestic Engi- 
neering 

Hospitals. 

Oak Park Infirmary, Electrical World 53: 912 

St. Luke's, Chicago, Power and Engineer 29: 996 



819 



Dec. 2, 1909 

May 23, 1908 
Apr. 13, 1909 

Sept. 12, 1908 

May 9, 1908 

Aug. 10, 1909 



Feb. 16, 1909 

May 16, 1908 
March 17, 1909 
Jan. 2, 1909 

May 30, 1908 
May 30, 1908 

May 10, 1910 

Nov. 30, 1909 

Feb. 13, 1909 

Feb. 22, 1908 

March 4, 1909 

Aug. 29, 1908 

April 19, 1910 

March 27, 1909 
Feb. 3, 1910 



July 17, 1909 
Dec. 25, 1909 



April 15, 1909 
Dec. 15, 1908 



820 STEAM POWER PLANT ENGINEERING 

Hotels. 

La Salle, Chicago, Electrical Review and West- 
ern Electrician 55: 418 Sept. 4, 1909 

Engineering Record 60: 241 Aug. 20, 1909 

New Plaza, New York, Engineering Record 58: 577 Nov. 21, 1908 

Miscellaneous. 

Carnegie Institute, Pittsburg, Power and Engi- 
neer 30: 97 Jan. 12, 1909 

West Point Military Academy, Power and Engi- 
neer 30: 747 April 27, 1909 

Manufacturing Plants. 

Allis-Chalmers Company, Milwaukee, Engineer- 
ing Record 61: 194 Feb. 12, 1910 

Corn Products Company, Argo, 111., Practical 

Engineer (United States) 14: 336 June, 1910 

Coronet Phosphate Company, Engineering 

Record 60: 469 Oct. 23, 1909 

Johnson & Johnson, New Brunswick, N. J., 

Power and Engineer 32: 568 March 29, 1910 

Heath & Milligan Paint Works, Engineering 

Record 59: 202 Feb. 20, 1909 

Sawmills, Power Plants for, Power 28: 973 June 23, 1908 

Textile Mills Power Plants, Electrical Review ... 53 : 638 Oct. 31, 1908 

Jan. 2, 1909 

Cassier's Magazine 34: 371 Aug., 1908 

Engineering Record 58: 440 Oct. 17, 1908 

Office Buildings. 

City Investing Building, New York, Power and 

Engineer 29: 983 Dec. 15, 1908 

Commercial National Bank, Chicago, Western 

Electrician 41 : 441 Dec. 7, 1907 

Kuper Building, Baltimore, Engineering 

Record 60: 492 Oct. 30, 1909 

Union National Bank Building, Pittsburg, 

Engineering Record 57: 818 June 27, 1908 

58: 13 July 4, 1908 

Power Plant Design. 

Central Electric Plants in Small Towns, Engi- 
neering News 62: 32 Sept. 23, 1909 

Central Stations vs. Private Plants, Engineering 87: 288 Feb. 26, 1909 

Central Stations, Towns over 1000 Inhabitants, 

Electrical Review, New York 56: 83 Jan. 8, 1910 

Design of Steam Electric Plants (Frank Koes- 

ter), Electrical Review (Serial) 54: 472 March 13, 190? 

From an Insurance Standpoint, Engineering 

Review 59: 295 March 13, 1909 

Heat Losses in an Electric Power Station, 

Engineering 87: 10 Jan. 1, 1909 



APPENDIX A 821 

Power Plant Design — Continued. 

Modern Power Station Design (H. B. Parsons), 

Cassier's Magazine 36: 328 Aug., 1909 

Prime Movers (C. P. Steinmetz), Pro. Am. 

Inst. Elec. Engrs 28: 135 Feb., 1909 

Progress in Power Plant Auxiliary Equipment, 

Street Railway Journal 31: 36 Jan. 11, 1908 

Simplicity in Steam Plant Design, Power 29: 371 Sept. 1, 1908 



APPENDIX B. 

RULES FOR CONDUCTING BOILER TRIALS * 
Code of 1899. 

I. Determine at the outset the specific object of the proposed trial, 
whether it be to ascertain the capacity of the boiler, its efficiency as a 
steam generator, its efficiency and its defects under usual working 
conditions, the economy of some particular kind of fuel, or the effect of 
changes of design, proportion, or operation; and prepare for the trial 
accordingly. (Appendix II.) 

II. Examine the boiler, both outside and inside; ascertain the dimen- 
sions of grates, heating surfaces, and all important parts; and make a 
full record, describing the same, and illustrating special features by 
sketches. The area of heating surface is to be computed from the sur- 
faces of shells, tubes, furnaces, and fire boxes in contact with the fire or 
hot gases. The outside diameter of water tubes and the inside diameter 
of fire tubes are to be used in the computation. All surfaces below the 
mean water level which have water on one side and products of com- 
bustion on the other are to be considered as water-heating surface, and 
all surfaces above the mean water level which have steam on one side 
and products of combustion on the other are to be considered as super- 
heating surface. 

III. Notice the general condition of the boiler and its equipment, and 
record such facts in relation thereto as bear upon the objects in view. 

If the object of the trial is to ascertain the maximum economy or 
capacity of the boiler as a steam generator, the boiler and all its appur- 
tenances should be put in first-class condition. Clean the heating surface 
inside and outside, remove clinkers from the grates and from the sides 
of the furnace. Remove all dust, soot, and ashes from the chambers, 
smoke connections, and flues. Close air leaks in the masonry and poorly 
fitted cleaning doors. See that the damper will open wide and close 
tight. Test for air leaks by firing a few shovels of smoky fuel and 

* From the report of the Committee of the American Society of Mechanical 
Engineers on the revision of the Society Code of 1885 relative to a standard method 
of conducting steam boiler trials. 

822 



APPENDIX B 823 

immediately closing the damper, observing the escape of smoke through 
the crevices, or by passing the flame of a candle over cracks in the 
brickwork. 

IV. Determine the character of the coal to be used. For tests of the 
efficiency or capacity of the boiler for comparison with other boilers the 
coal should, if possible, be of some kind which is commercially regarded 
as a standard. For New England and that portion of the country east 
of the Allegheny Mountains, good anthracite egg coal, containing not 
over 10 per cent of ash, and semi-bituminous Clearfield (Pa.), Cumber- 
land (Md.), and Pocahontas (Va.) coals are thus regarded. West of the 
Allegheny Mountains, Pocahontas (Va.) and New River (W. Va.) 
semi-bituminous, and Youghiogheny or Pittsburg bituminous coals are 
recognized as standards.* There is no special grade of coal mined in 
the Western States which is widely recognized as of superior quality 
or considered as a standard coal for boiler testing. Big Muddy lump, an 
Illinois coal mined in Jackson County, 111., is suggested as being of 
sufficiently high grade to answer the requirements in districts where 
it is more conveniently obtainable than the other coals mentioned 
above. 

For tests made to determine the performance of a boiler with a par- 
ticular kind of coal, such as may be specified in a contract for the sale 
of a boiler, the coal used should not be higher in ash and in moisture 
than that specified, since increase in ash and moisture above a stated 
amount is apt to cause a falling off of both capacity and economy in 
greater proportion than the proportion of such increase. 

V. Establish the correctness of all apparatus used in the test for weigh- 
ing and measuring. These are: 

1. Scales for weighing coal, ashes, and water. 

2. Tanks, or water meters for measuring water. Water meters, as 
a rule, should only be used as a check on other measurements. For 
accurate work, the water should be weighed or measured in a tank. 
(Appendices I, IV, VII, VIII.) 

3. Thermometers and pyrometers for taking temperatures of air, 
steam, feed water, waste gases, etc. (Appendix XXVII.) 

4. Pressure gauges, draught gauges, etc. (Appendices XXVIII to 
XXX.) 

* These coals are selected because they are about the only coals which contain 
the essentials of excellence of quality, adaptability to various kinds of furnaces, 
grates, boilers, and methods of firing, and wide distribution and general accessibility 
in the markets. 



824 STEAM POWER PLANT ENGINEERING 

The kind and location of the various pieces of testing apparatus must 
be left to the judgment of the person conducting the test; always keeping 
in mind the main object, i.e., to obtain authentic data. 

VI. See that the boiler is thoroughly heated before the trial to its usual 
working temperature. If the boiler is new and of a form provided with 
a brick setting, it should be in regular use at least a week before the 
trial, so as to dry and heat the walls. If it has been laid off and become 
cold, it should be worked before the trial until the walls are well heated. 

VII. The boiler and connections should be proved to be free from 
leaks before beginning a test, and all water connections, including blow 
and extra feed pipes, should be disconnected, stopped with blank 
flanges, or bled through special openings beyond the valves, except the 
particular pipe through which water is to be fed to the boiler during the 
trial. During the test the blow-off and feed pipes should remain exposed 
to view. 

If an injector is used, it should receive steam directly through a felted 
pipe from the boiler being tested.* 

If the water is metered after it passes the injector, its temperature 
should be taken at the point where it leaves the injector. If the quan- 
tity is determined before it goes to the injector the temperature should 
be determined on the suction side of the injector, and if no change of 
temperature occurs other than that due to the injector, the tempera- 
ture thus determined is properly that of the feed water. When the 
temperature changes between the injector and the boiler, as by the use 
of a heater or by radiation, the temperature at which the water enters and 
leaves the injector and that at which it enters the boiler should all be 
taken. In that case the weight to be used is that of the water leaving 
the injector, computed from the heat units if not directly measured, 
and the temperature that of the water entering the boiler. 

Let w = weight of water entering the injector. 
x = weight of steam entering the injector. 
h t = heat units per pound of water entering injector. 
h 2 = heat units per pound of steam entering injector. 
h 3 = heat units per pound of water leaving injector. 

* In feeding a boiler undergoing test with an injector taking steam from another 
boiler, or from the main steam pipe from several boilers, the evaporative results 
may be modified by a difference in the quality of the steam from such source com- 
pared with that supplied by the boiler being tested, and in some cases the connection 
to the injector may act as a drip for the main steam pipe. If it is known that the 
steam from the main pipe is of the same pressure and quality as that furnished by 
the boiler undergoing the test, the steam may be taken from such main pipe. 



APPENDIX B 825 



Then w + x = weight of water leaving injector. 

h» — h-. 
x = w -r 3 L « 

h 2 — h a 

See that the steam main is so arranged that water of condensation 
cannot run back into the boiler. 

VIII. Duration of the Test. — For tests made to ascertain either the 
maximum economy or the maximum capacity of a boiler, irrespective 
of the particular class of service for which it is regularly used, the 
duration should be at least 10 hours of continuous running. If the 
rate of combustion exceeds 25 pounds of coal per square foot of grate 
surface per hour, it may be stopped when a total of 250 pounds of coal 
has been burned per square foot of grate. 

In cases where the service requires continuous running for the 
whole 24 hours of the day, with shifts of firemen a number of times 
during that period, it is well to continue the test for at least 24 
hours. 

When it is desired to ascertain the performance under the working 
conditions of practical running, whether the boiler be regularly in use 
24 hours a day or only a certain number of hours out of each 24, the 
fires being banked the balance of the time, the duration should not be 
less than 24 hours. 

IX. Starting and Stopping a Test. — The conditions of the boiler 
and furnace in all respects should be, as nearly as possible, the same at 
the end as at the beginning of the test. The steam pressure should be 
the same; the water level the same; the fire upon the grates should 
be the same in quantity and condition; and the walls, flues, etc., 
should be of the same temperature. Two methods of obtaining the 
desired equality of conditions of the fire may be used, viz.: those 
which were called in the Code of 1885 " the standard method " and 
" the alternate method," the latter being employed where it is incon- 
venient to make use of the standard method.* 

X. Standard Method of Starting and Stopping a Test. — Steam being 
raised to the working pressure, remove rapidly all the fire from the 
grate, close the damper, clean the ash pit, and as quickly as possible 
start a new fire with weighed wood and coal, noting the time and the 

* The Committee concludes that it is best to retain the designations " standard " 
and "alternate," since they have become widely known and established in the 
minds of engineers and in the reprints of the Code of 1885. Many engineers prefer 
the "alternate" to the "standard" method on account of its being less liable to 
error due to cooling of the boiler at the beginning and end of a test. 



£26 STEAM POWER PLANT ENGINEERING 

water level* while the water is in a quiescent state, just before lighting 
the fire. 

At the end of the test remove the whole fire, which has been burned 
low, clean the grates and ash pit, and note the water level when the 
water is in a quiescent state, and record the time of hauling the fire. 
The water level should be as nearly as possible the same as at the 
beginning of the test. If it is not the same, a correction should be 
made by computation, and not by operating the pump after the test is 
completed. 

XI. Alternate Method of Starting and Stopping a Test. — The boiler 
being thoroughly heated by a preliminary run, the fires are to be 
burned low and well cleaned. Note the amount of coal left on the 
grate as nearly as it can be estimated; note the pressure of steam and 
the water level. Note the time, and record it as the starting time. 
Fresh coal which has been weighed should now be fired. The ash pits 
should be thoroughly cleaned at once after starting. Before the end 
of the test the fires should be burned low, just as before the start, and 
the fires cleaned in such a manner as to leave a bed of coal on the 
grates of the same depth, and in the same condition, as at the start. 
When this stage is reached, note the time and record it as the stopping 
time. The water level and steam pressures should previously be 
brought as nearly as possible to the same point as at the start. If the 
water level is not the same as at the start, a correction should be 
made by computation, and not by operating the pump after the test 
is completed. 

XII. Uniformity of Conditions. — In all trials made to ascertain 
maximum economy or capacity, the conditions should be maintained 
uniformly constant. Arrangements should be made to dispose of the 
steam so that the rate of evaporation may be kept the same from 
beginning to end. This may be accomplished in a single boiler by 
carrying the steam through a waste steam pipe, the discharge from 
which can be regulated as desired. In a battery of boilers, in which 
only one is tested, the draft can be regulated on the remaining 
boilers, leaving the test boiler to work under a constant rate of 
production. 

Uniformity of conditions should prevail as to the pressure of steam, 
the height of water, the rate of evaporation, the thickness of fire, the 

* The gauge glass should not be blown out within an hour before the water 
level is taken at the beginning and end of a test, otherwise an error in the reading 
of the water level may be caused by a change in the temperature and density of the 
water in the pipe leading from the bottom of the glass into the boiler. 



APPENDIX B 827 

times of firing and quantity of coal fired at one time, and as to the 
intervals between the times of cleaning the fires. 

The method of firing to be carried on in such tests should be dic- 
tated by the expert or person in responsible charge of the test, and 
the method adopted should be adhered to by the fireman throughout 
the test. 

XIII. Keeping the Records. — Take note of every event connected 
with the progress of the trial, however unimportant it may appear. 
Record the time of every occurrence and the time of taking every 
weight and every observation. (Appendices I, IV, V, VI, VII, .VIII.) 

The coal should be weighed and delivered to the fireman in equal 
proportions, each sufficient for not more than one hour's run, and a 
fresh portion should not be delivered until the previous one has all 
been fired. The time required to consume each portion should be 
noted, the time being recorded at the instant of firing the last of each 
portion. It is desirable that at the same time the amount of water 
fed into the boiler should be accurately noted and recorded, including 
the height of the water in the boiler and the average pressure of 
steam and temperature of feed during the time. By thus recording 
the amount of water evaporated by successive portions of coal, the 
test may be divided into several periods if desired, and the degree 
of uniformity of combustion, evaporation, and economy analyzed for 
each period. In addition to these records of the coal and the feed 
water, half-hourly observations should be made of the temperature of 
the feed water, of the flue gases, of the external air in the boiler room, 
of the temperature of the furnace when a furnace pyrometer is used, 
also of the pressure of steam, and of the readings of the instruments 
for determining the moisture in the steam. A log should be kept on 
properly prepared blanks containing columns for record of the various 
observations. (Appendix XXII.) 

When the " standard method ".of starting and stopping the test is 
used, the hourly rate of combustion and of evaporation and the horse 
power should be computed from the records taken during the time 
when the fires are in active condition. This time is somewhat less 
than the actual time which elapses between the beginning and end of 
the run. The loss of time due to kindling the fire at the beginning and 
burning it out at the end makes this course necessary. 

XIV. Quality of Steam. — The percentage of moisture in the steam 
should be determined by the use of either a throttling or a separating 
steam calorimeter. The sampling nozzle should be placed in the ver- 
tical steam pipe rising from the boiler. It should be made of J-inch 



828 STEAM POWER PLANT ENGINEERING 

pipe, and should extend across the diameter of the steam pipe to 
within half an inch of the opposite side, being closed at the end and 
perforated with not less than twenty J-inch holes equally distributed 
along and around its cylindrical surface, but none of these holes should 
be nearer than \ inch to the inner side of the steam pipe. The calo- 
rimeter and the pipe leading to it should be well covered with felting. 
Whenever the indications of the throttling or separating calorimeter 
show that the percentage of moisture is irregular, or occasionally in 
excess of three per cent, the results should be checked by a steam 
separator placed in the steam pipe as close to the boiler as convenient, 
with a calorimeter in the steam pipe just beyond the outlet from 
the separator. The drip from the separator should be caught 
and weighed, and the percentage of moisture computed therefrom 
added to that shown by the calorimeter. (See Appendices XV to 
XVII.) 

Superheating should be determined by means of a thermometer 
placed in a mercury well inserted in the steam pipe. The degree of 
superheating should be taken as the difference between the reading 
of the thermometer for superheated steam and the readings of the 
same thermometer for saturated steam at the same pressure as 
determined by a special experiment, and not by reference to steam 
tables. 

For calculations relating to quality of steam and corrections for 
quality of steam, see Appendices XVIII and XIX. 

XV. Sampling the Coal and Determining its Moisture. — As each 
barrow load or fresh portion of coal is taken from the coal pile, a rep- 
resentative shovelful is selected from it and placed in a barrel or box 
in a cool place and kept until the end of the trial. The samples are 
then mixed and broken into pieces not exceeding one inch in diameter, 
and reduced by the process of repeated quartering and crushing until 
a final sample weighing about five pounds is obtained, and the sizes of 
the larger pieces are such that they will pass through a sieve with 
J-inch meshes. From this sample two one-quart, air-tight glass pre- 
serving jars, or other air-tight vessels which will prevent the escape of 
moisture from the sample, are to be promptly filled, and these samples 
are to be kept for subsequent determinations of moisture and of heat- 
ing value and for chemical analyses. During the process of quar- 
tering, when the sample has been reduced to about 100 pounds, a 
quarter to a half of it may be taken for an approximate determination 
of moisture. This may be made by placing it in a shallow iron pan, 
not over three inches deep, carefully weighing it, and setting the pan 
in the hottest place that can be found on the brickwork of the boiler 



APPENDIX B 829 

setting or flues, keeping it there for at least 12 hours, and then 
weighing it. The determination of moisture thus made is believed 
to be approximately accurate for anthracite and semi-bituminous 
coals, and also for Pittsburg or Youghiogheny coal; but it cannot be 
relied upon for coals mined west of Pittsburg, or for other coals con- 
taining inherent moisture. For these latter coals it is important that 
a more accurate method be adopted. The method recommended by 
the Committee for all accurate tests, whatever the character of the 
coal, is described as follows: 

Take one of the samples contained in the glass jars and subject it to 
a thorough air-drying by spreading it in a thin layer and exposing it 
for several hours to the atmosphere of a warm room, weighing it 
before and after, thereby determining the quantity of surface moisture 
it contains. Then crush the whole of it by running it through an 
ordinary coffee mill adjusted so as to produce somewhat coarse grains 
(less than T V inch), thoroughly mix the crushed sample, select from it 
a portion of from 10 to 50 grams, weigh it in a balance which will 
easily show a variation as small as 1 part in 1000, and dry it in an air 
or sand bath at a temperature between 240 and 280 degrees F. for one 
hour. Weigh it and record the loss, then heat and weigh it again 
repeatedly, at intervals of an hour or less, until the minimum weight 
has been reached and the weight begins to increase by oxidation of 
a portion of the coal. The difference between the original and the 
minimum weight is taken as the moisture in the air-dried coal. This 
moisture test should preferably be made on duplicate samples, and the 
results should agree within 0.3 to 0.4 of one per cent, the mean of 
the two determinations being taken as the correct result. The sum of 
the percentage of moisture thus found and the percentage of surface 
moisture previously determined is the total moisture. (Appendix XL) 

XVI. Treatment of Ashes and Refuse. — The ashes and refuse are 
to be weighed in a dry state. If it is found desirable to show the 
principal characteristics of the ash, a sample should be subjected to a 
proximate analysis and the actual amount of incombustible material 
determined. For elaborate trials a complete analysis of the ash and 
refuse should be made. 

XVII. Calorific Tests and Analysis of Coal. — The quality of the 
fuel should be determined either by heat test or by analysis, or by both. 

The rational method of determining the total heat of combustion is 
to burn the sample of coal in an atmosphere of oxygen gas, the coal to 
be sampled as directed in Article XV of this code. (See Appendices 
XIII and XIV.) 



830 STEAM POWER PLANT ENGINEERING 

The chemical analysis of the coal should be made only by an 
expert chemist. The total heat of combustion computed from the 
results of the ultimate analysis may be obtained by the use of Dulong's 
formula (with constants modified by recent determinations), viz.: 

14,600 C + 62,000 (h -Q-\ + 4000 S, in which C, H, O, and S refer 

to the proportions of carbon, hydrogen, oxygen, and sulphur respec- 
tively, as determined by the ultimate analysis.* 

It is desirable that a proximate analysis should be made, thereby 
determining the relative proportions of volatile matter and fixed 
carbon. These proportions furnish an indication of the leading 
characteristics of the fuel, and serve to fix the class to which it 
belongs. (Appendix XII.) As an additional indication of the char- 
acteristics of the fuel, the specific gravity should be determined. 

XVIII. Analysis of Flue Gases. — The analysis of the flue gases is 
an especially valuable method of determining the relative value of 
different methods of firing, or of different kinds of furnaces. In 
making these analyses great care should be taken to procure average 
samples, since the composition is apt to vary at different points of 
the flue. (Appendix XXXI.) The composition is also apt to vary 
from minute to minute, and for this reason the drawings of gas 
should last a considerable period of time. Where complete de- 
terminations are desired, the analyses should be intrusted to an 
expert chemist. For approximate determinations the Orsatf or the 
Hempel | apparatus may be used by the engineer. (See Appendix 
XXXIII.) 

For the continuous indication of the amount of carbonic acid 
present in the flue gases, an instrument may be employed which 
shows the weight of the sample of gas passing through it. (Appendix 
XXXIX.) 

XIX. Smoke Observations. — It is desirable to have a uniform system 
of determining and recording the quantity of smoke produced where 
bituminous coal is used. The system commonly employed is to express 
the degree of smokiness by means of percentages dependent upon the 
judgment of the observer. The Committee does not place much value 

* Favre and Silberman give 14,544 B.T.U. per pound carbon; Berthelot 14,647 
B.T.U. Favre and Silberman give 62,032 B.T.U. per pound hydrogen; Thomsen 
61,816 B.T.U. 

f See R. S. Hale's paper on "Flue Gas Analysis," Transactions, Vol. XVIII, 
p. 901. 

t See Hempel on " Gas Analysis. " 



APPENDIX B 831 

upon a percentage method, because it depends so largely upon the per- 
sonal element, but if this method is used, it is desirable that, so far as 
possible, a definition be given in explicit terms as to the basis and method 
employed in arriving at the percentage. The actual measurement of a 
sample of soot and smoke by some form of meter is to be preferred. 
(See Appendices XXXIV and XXXV.) 

XX. Miscellaneous. — In tests for purposes of scientific research, in 
which the determination of all the variables entering into the test is 
desired, certain observations should be made which are in general 
unnecessary for ordinary tests. These are the measurement of the air 
supply, the determination of its contained moisture, the determination 
of the amount of heat lost by radiation, of the amount of infiltration of 
air through the setting, and (by condensation of all the steam made by 
the boiler) of the total heat imparted to the water. 

As these determinations are rarely undertaken, it is not deemed 
advisable to give directions for making them. 

XXI. Calculations of Efficiency. — Two methods of defining and 
calculating the efficiency of a boiler are recommended. They are: 

i ~nrc - £ j-u u -i Heat absorbed per lb. combustible 

1. Efficiency of the boiler — -— — — ■ ^ — - — - • 

Calorific value of 1 lb. combustible 

2. Efficiency of the boiler and grate = Heat absorbed per lb coal . 

Calorific value of 1 lb. coal 

The first of these is sometimes called the efficiency based on com- 
bustible, and the second the efficiency based on coal. The first is 
recommended as a standard of comparison for all tests, and this is the 
one which is understood to be referred to when the word " efficiency " 
alone is used without qualification. The second, however, should be 
included in a report of a test, together with the first, whenever the 
object of the test is to determine the efficiency of the boiler and furnace 
together with the grate (or mechanical stoker), or to compare different 
furnaces, grates, fuels, or methods of firing. 

The heat absorbed per pound of combustible (or per pound coal) is 
to be calculated by multiplying the equivalent evaporation from and 
at 212 degrees per pound combustible (or coal) by 965.7 (Appendix 
XX.) 

XXII. The Heat Balance. — An approximate " heat balance," or 
statement of the distribution of the heating value of the coal among the 
several items of heat utilized and heat lost, may be included in the 



832 



STEAM POWER PLANT ENGINEERING 



report of a test when analyses of the fuel and of the chimney gases have 
been made. It should be reported in the following form: 

HEAT BALANCE, OR DISTRIBUTION OF THE HEATING VALUE OF THE 

COMBUSTIBLE. 

Total Heat Value of 1 pound of combustible B.T.U. 



3. 



5. 



Heat absorbed by the boiler = evaporation from and at 212 

degrees per pound of combustible X 965.7. 
Loss due to moisture in coal = per cent of moisture referred 

to combustible -> 100 X [(212 - t) + 966 + 0.48 (T - 

212)] (l = temperature of air in the boiler room, T = 

that of the flue gases). 
Loss due to moisture formed by the burning of hydrogen 

= per cent of hydrogen to combustible -f- 100 X 9 X 

[(212 - t) + 966 + 0.48 (T - 212)]. 
* Loss due to heat carried away in the dry chimney gases = 

weight of gas per pound of combustible X 0.24 X 

(T - t). 

CO 
f Loss due to incomplete combustion of carbon = 



per cent C in combustible , 
X — X x0 ; 150. 



C0 2 +CO 



Loss due to unconsumed hydrogen and hydrocarbons, to 
heating the moisture in the air, to radiation, and un- 
accounted for. (Some of these losses may be sepa- 
rately itemized if data are obtained from which they 
may be calculated.) 



Totals. 



Per Cent. 




* The weight of gas per pound of carbon burned may be calculated from the gas analyses as 

follows: 

11 C0 2 +80+7 (CO + N) 

Dry gas per pound carbon = ■ . in which C0 2 , CO, O, and N are 

3 (COo + CO) 



the percentages by volume of the several gases. As the sampling and analyses of the gases in the 
present state of the art are liable to considerable errors, the result of this calculation is usually 
only an approximate one. The heat balance itself is also only approximate for this reason, as well 
as for the fact that it is not possible to determine accurately the percentage of unburned hydrogen 
or hydrocarbons in the flue gases. (See Appendix XXXII.) 

The weight of dry gas per pound of combustible is found by multiplying the dry gas per pound 
of carbon by the percentage of carbon in the combustible and dividing by 100. 

f C0 2 and CO are respectively the percentage by volume of carbonic acid and carbonic oxide in 
the flue gases. The quantity 10,150 = No. heat units generated by burning to carbonic acid 
one pound of carbon contained in carbonic oxide. 

XXIII. Report of the Trial. — The data and results should be 
reported in the manner given in either one of the two following tables, 
omitting lines where the tests have not been made as elaborately as 
provided for in such tables. Additional lines may be added for data 
relating to the specific object of the test. The extra lines should be 



APPENDIX B 833 

classified under the headings provided in the tables, and numbered as 
per preceding line, with sub letters a, b, etc. The Short Form of Report, 
Table No. 2, is recommended for commercial tests and as a convenient 
form of abridging the longer form for publication when saving of space 
is desirable. For elaborate trials, it is recommended that the full log 
of the trial be shown graphically, by means of a chart. (Appendix 
XXXVIII.) 

TABLE NO. 1. 
Data and Results of Evaporative Test, 

Arranged in accordance with the Complete Form advised by the Boiler Test Com- 
mittee of the American Society of Mechanical Engineers. Code of 1899. 

Made by of boiler at ■ to 

determine 

Principal conditions governing the trial 



Kind of fuel * 

Kind of furnace 

State of the weather . 



Method of starting and stopping the test ("standard" or "alternate," Art. X 
and XI, Code) 

1. Date of trial . 

2. Duration of trial hours. 

Dimensions and Proportions. 

A complete description of the boiler, and drawings of the same if of unusual 
type, should be given on an annexed sheet. (See Appendix X.) 

3. Grate surface width length area square feet 

4. Height of furnace inches. 

5. Approximate width of air spaces in grate inch. 

6. Proportion of air space to whole grate surface per cent. 

7. Water-heating surface square feet. 

8. Superheating surface " 

9. Ratio of water-heating surface to grate surface — to 1. 

10. Ratio of minimum draft area to grate surface 1 to — . 

Average Pressures. 

11. Steam pressure by gauge lb. per sq. in. 

12. Force of draft between damper and boiler in. of water. 

13. Force of draft in furnace " " 

14. Force of draft or blast in ash pit " " 

* The items printed in italics correspond to the items in the " Short Form of 
Code." 



834 STEAM POWER PLANT ENGINEERING 

Average Temperatures. 

15. Of external air degrees. 

16. Of fire room " 

17. Of steam " 

18. Of feed water entering heater " 

19. Of feed water entering economizer " 

20. Of feed water entering boiler " 

21. Of escaping gases from boiler " 

22. Of escaping gases from economizer " 

Fuel. 

23. Size and condition 

24. Weight of wood used in lighting fire pounds. 

25. Weight of coal as fired * , " 

26. Percentage of moisture in coal f per cent. 

27. Total weight of dry coal consumed pounds. 

28. Total ash and refuse 

29. Quality of ash and refuse 

30. Total combustible consumed pounds. 

31. Percentage of ash and refuse in dry coal per cent. 

Proximate Analysis of Coal. 

(App. XII.) 

Of Coal. Of Combustible. 

32. Fixed carbon per cent. per cent. 

33 Volatile matter 

34. Moisture " 

35. Ash " 

100 per cent. 100 per cent. 

36. Sulphur, separately determined 

Ultimate Analysis of Dry Coal. 

(Art. XVII, Code.) 

Of Coal. Of Combustible. 

37. Carbon (C) per cent. per cent. 

38. Hydrogen (H) 

39. Oxygen (O) 

40. Nitrogen (N) 

41. Sulphur(S) 

42. Ash " — 

100 per cent. 100 per cent. 

43. Moisture in sample of coal as received 

* Including equivalent of wood used in lighting the fire, not including unburned 
coal withdrawn from furnace at times of cleaning and at end of test. One pound of 
wood is taken to be equal to 0.4 pound of coal, or, in case greater accuracy is 
desired, as having a heat value equivalent to the evaporation of 6 pounds of water 
from and at 212 degrees per pound. (6 X 965.7 = 5794 B.T.U.) 

f This is the total moisture in the coal as found by drying it artificially, as 
described in Art. XV of Code. 



APPENDIX B 835 

Analysis of Ash and Refuse. 

44. Carbon per cent. 

45. Earthy matter " 

Fuel per Hour. 

46. Dry coal consumed per hour pounds. 

47. Combustible consumed per hour 

. 48. Dry coal per square foot of grate surface per hour 

49. Combustible per square foot of water-heating surface per hour. . . 

Calorific Value of Fuel. 
(Art. XVII, Code.) 

50. Calorific value by oxygen calorimeter, per pound of dry coal B.T.U. 

51. Calorific value by oxygen calorimeter, per pound of combustible 

52. Calorific value by analysis, per pound of dry coal * 

53. Calorific value by analysis, per pound of combustible 

Quality of Steam. 
(App. XV to XIX.) 

54. Percentage of moisture in steam per cent. 

55. Number of degrees of superheating degrees. 

56. Quality of steam (dry steam = unity). (For exact determination 

of the factor of correction for quality of steam see Appendix 
XVIII) 

Water. 
(App. I, IV, VII, VIII.) 

57. Total weight of water fed to boiler f pounds. 

58. Equivalent water fed to boiler from and at 212 degrees 

59. Water actually evaporated, corrected for quality of steam 

60. Factor of evaporation J pounds. 

61. Equivalent water evaporated into dry steam from and at 212 

degrees. (Item 59 X Item 60.) " 

Water per Hour. 

62. Water evaporated per hour, corrected for quality of steam 

63. Equivalent evaporation per hour from and at 212 degrees 

64. Equivalent evaporation per hour from and at 212 degrees per square 

foot of water-heating surface 

* See formula for calorific value under Article XVII of Code, 
f Corrected for inequality of water level and of steam pressure at beginning and 
end of test. 

t Factor of evaporation = ,__ _ in which H and h are respectively the total heat 
965.7 

in steam of the average observed pressure, and in water of the average observed 

temperature of the feed. 



836 



STEAM POWER PLANT ENGINEERING 



Horse Power. 

65. Horse power developed. (34| pounds of water evaporated per hour 

into dry steam from and at 212 degrees equals one horse power.)* ... horse power. 

66. Builders' rated horse power " 

67. Percentage of builders' rated horse power developed per cent. 



Economic Results. 

68. Water apparently evaporated per pound of coal under actual condi- 

tions. (Item 58 -s- Item 25.) 

69. Equivalent evaporation from and at 212 degrees per pound of wet 

coal. (Item 61 -f- Item 25.) 

70. Equivalent evaporation from and at 212 degrees per pound of dry 

coal. (Item 61 -j- Item 27.) 

71. Equivalent evaporation from and at 212 degrees per pound of com- 

bustible. (Item 61 -f- Item 30.) 

(If the equivalent evaporation, Items 69, 70, and 71, is not cor- 
rected for the quality of steam, the fact should be stated.) 



pounds. 



Efficiency. 
(See Art. XXI, Code.) 

72. Efficiency of the boiler; heat absorbed by the boiler per pound of com- 

bustible divided by the heat value of one pound of combustible f . . . per cent. 

73. Efficiency of boiler, including the grate; heat absorbed by the boiler, 

per pound of dry coal fired, divided by the heat value of one pound 

of dry coal X " 

Cost of Evaporation. 

74. Cost of coal per ton of pounds delivered in boiler room $ 

75. Cost of fuel for evaporating 1000 pounds of water under observed 

conditions $ 

76. Cost of fuel used for evaporating 1000 pounds of water from and at 

212 degrees $ 

Smoke Observations, 
(App. XXXIV and XXXV.) 

77. Percentage of smoke as observed per cent. 

78. Weight of soot per hour obtained from smoke meter ounces. 

79. Volume of soot per hour obtained from smoke meter cubic inches. 

* Held to be the equivalent of 30 pounds of water per hour evaporated from 100 
degrees F. into dry steam at 70 pounds gauge pressure. (See Introduction to Code.) 

f In all cases where the word combustible is used, it means the coal without 
moisture and ash, but including all other constituents. It is the same as what is 
called in Europe "coal dry and free from ash." 

J The heat value of the coal is to be determined either by an oxygen calorimeter 
or by calculation from ultimate analysis. 



APPENDIX B 



837 



Methods of Firing. 

80. Kind of firing (spreading, alternate, or coking) 

81. Average thickness of fire 

82. Average intervals between firings for each furnace during time 

when fires are in normal condition 

83. Average interval between times of levelling or breaking up 

Analyses of the Dry Gases. 

84. Carbon dioxide (C0 2 ) per cent. 

85. Oxygen (O) 

86. Carbon monoxide (CO) " 

87. Hydrogen and hydrocarbons " 

88. Nitrogen (by difference) (N) 

100 per cent. 



TABLE NO. 2. 
Data and Results of Evaporative Test, 

Arranged in accordance with the Short Form advised by the Boiler Test Committee 
of the American Society of Mechanical Engineers. Code of 1899. 

Made by on boiler, at to 

determine 

Kind of fuel 

Kind of furnace 

Method of starting and stopping the test (" standard" or "alternate," Art. X 

and XI, Code) 

Grate surface square feet. 

Water-heating surface " 

Superheating surface " 



Total 

1. Date of trial 

2. Duration of trial hours. 

3. Weight of coal as fired * pounds. 

4. Percentage of moisture in coal * per cent. 

5. Total weight of dry coal consumed pounds. 

6. Total ash and refuse " 

7. Percentage of ash and refuse in dry coal per cent. 

8. Total weight of water fed to the boiler pounds. 

9. Water actually evaporated, corrected for moisture or superheat in 

• steam " 

10. Equivalent water evaporated into dry steam from and at 212 



* See foot-notes of Complete Form. 



838 STEAM POWER PLANT ENGINEERING 



Hourly Quantities. 

11. Dry coal consumed per hour pounds. 

12. Dry coal per hour per square foot of grate surface " 

13. Water fed per hour " 

14. Equivalent water evaporated per hour from and at 212 degrees 

corrected for quality of steam " 

15. Equivalent water evaporated per hour per square foot of water- 

heating surface " 



Average Pressures, Temperatures, etc. 

16. Average boiler pressure lb. per sq. in. 

17. Average temperature of feed water degrees. 

18. Average temperature of escaping gases " 

19. Average force of draft between damper and boiler in. of water. 

20. Percentage of moisture in steam, or number of degrees of super- 

heating 

Horse Power. 

21. Horse power developed (Item 14 -f- 34^) horse power. 

22. Builders' rated horse power " 

23. Percentage of builders' rated horse power developed per cent. 

Economic Results. 

24. Water apparently evaporated per pound of coal under actual 

conditions. (Item 8 -5- Item 3.) pounds. 

25. Equivalent water actually evaporated from and at 212 degrees 

per pound of wet coal. (Item 9 -r- Item 3.) " 

26. Equivalent evaporation from and at 212 degrees per pound of dry 

coal. (Item 9 -f- Item 5.) " 

27. Equivalent evaporation from and at 212 degrees per pound of 

combustible. [Item 9 -*- (Item 5 — Item 6).] " 

(If Items 25, 26, and 27 are not corrected for quality of steam, the 
fact should be stated.) 

Efficiency. 

28. Heating value of the coal per pound B.T.IL 

29. Heating value of the combustible per pound " 

30. Efficiency of boiler (based on combustible) per cent. 

31. Efficiency of "boiler, including grate (based on coal) " 

Cost of Evaporation. 

32. Cost of coal per ton delivered in boiler room $ 

33. Cost of coal required for evaporation of 1000 pounds of water 

from and at 212 degrees $ 



APPENDIX B 839 

LIST OF APPENDICES TO CODE * 

No. of Appendix. 

I. Relative Weights of Water and Fuel c. e. e. 

II. Object of the Test. (I, 1885 Code) j. c. h. 

III. General Observations. (II, 1885 Code) c. t. p. 

IV. Precautions to be Observed in Making a Boiler Test. (Ill, 1885 

Code) c. e. e. 

V. Weighing the Coal. (IV, 1885 Code) j. c. h. 

VI. Weighing the Coal. (V, 1885 Code) c. t. p. 

VII. Weighing the Water. (VI, 1885 Code) j. c. h. 

VIII. Measuring the Feed Water. (VII, 1885 Code) c. t. p. 

IX. Keeping Time of Observations. (VIII, 1885 Code) j. c. h. 

X. Description of Boiler. (XXIII, 1885 Code) . . .c. e. e. 

XI. Determining the Moisture in Coal w. k. 

XII. Proximate Analyses of Coal . . w. k. 

XIII. Coal Calorimeter g. h. b. 

XIV. Comparative Calorimetric Tests of Coals w. k. 

XV. Determination of the Moisture in the Steam w. k. 

XVI. Correction for Radiation from Throttling Calorimeters g. h. b. 

XVII. Combined Calorimeter and Separator g. h. b. 

XVIII. Corrections for Quality of Steam c. e. e. 

XIX. The Quality of Superheated Steam g. h. b. 

XX. Efficiency of the Boiler w. k. 

XXI. Distribution of the Heating Value of the Fuel . . w. k. 

XXII. Observation Blanks. (Amendment to XXIV, 1885 Code) . . . .c. e. e. 

XXIII. Horse Power. (XXV, 1885 Code) j. c. h. 

XXIV. Steam Units. (XXVI, 1885 Code) c. e. e. 

XXV. Discrepancy between Commercial and Experimental Results. .. .c. e. e. 

XXVI. Recording Steam Gauge. (IX, 1885 Code) j. c. h. 

XXVII. Pyrometer. (XIII, 1885 Code) c. t. p. 

XXVIII. Draft Gauge. (XIV, 1885 Code) j. c. h. 

XXIX. Draft Gauge g. h. b. 

XXX. Draft Gauge w. k. 

XXXI. Sampling Flue Gases. (XVI, 1885 Code) j. c. h. 

XXXII. Computation of the Weight of Chimney Gases from the Analysis 

by Volume of Dry Gas w. k. 

XXXIII. The Orsat Apparatus for Analyzing Flue Gases g. h. b. 

XXXIV. Smoke Measurements g. h. b. 

XXXV. The Ringelmann Smoke Chart w. k. 

XXXVI. Starting and Stopping a Test w. k. 

XXXVII. Starting and Stopping a Test g. h. b. 

XXXVIII. Chart Showing Graphically the Log of a Trial g. h. b. 

XXXIX. Continuous Determinations of Carbonic Acid in Flue Gases. . . .g. h. b. 

XL.fMeasuring Radiation from Certain Types of Boilers .r. s. h. 

XLI.f Determination of the Moisture in Steam Flowing through a 

Horizontal Pipe d. s. j. 

* Only a few of the appendices are reprinted. 

t Contributed by members of the Society and accepted by the Committee for 
publication in the Appendix. 



840 STEAM POWER PLANT ENGINEERING 

APPENDIX XX. 
Efficiency of the Boiler. 

The efficiency of the boiler, including the grate, or the efficiency based 
on coal, is the quotient arising from dividing the heat absorbed by the 
boiler by the heating value of the total amount of coal supplied to the 
boiler, including the coal which falls through the grate. It may be 
conveniently calculated by multiplying the number of pounds of water 
evaporated from and at 212 degrees F. into dry steam per pound of 
dry coal by 965.7 and dividing the product by the heating value in 
British thermal units of one pound of dry coal. 

The efficiency of the boiler, not including the grate, or the efficiency 
based on combustible, is the quotient arising from dividing the heat 
absorbed by the boiler by the heating value of the combustible burned. 
It may be calculated by multiplying the number of pounds of water 
evaporated from and at 212 degrees F. into dry steam per pound of 
combustible by 965.7 and dividing the product by the heating value in 
British thermal units of one pound of combustible; the term " com- 
bustible " being denned as coal dry and free from ash, or the coal sup- 
plied to the boiler less its moisture and the ash and unburned coal 
which falls through the grate or is otherwise withdrawn from the furnace. 

The efficiency of the boiler, not including the grate (or the efficiency 
based upon combustible), is a more accurate measure of comparison of 
different boilers than the efficiency including the grate (or the efficiency 
based upon coal), for the latter is subject to a number of variable con- 
ditions, such as size and character of the coal, air spaces between the 
grate bars, skill of the fireman in saving coal from falling through the 
grate, etc. It is, moreover, subject to errors of sampling the coal for 
drying and for analysis, which affect the result to a greater degree than 
they do the efficiency based upon combustible, for the reason that the 
heating value of one pound of combustible of any sample selected from 
a given lot, such as a car load, of coal is practically a constant quantity 
and is independent of the percentage of moisture and ash in the sample; 
while the sample itself, upon the heating value of which the efficiency 
based on coal is calculated, may differ in its percentage of moisture and 
ash from the average coal used in the boiler test. 

When the object of a boiler test is to determine its efficiency as an 
absorber of heat, or to compare it with other boilers, the efficiency based 
on combustible is the one which should be used; but when the object of 
the test is to determine the efficiency of the combination of the boiler, 
the furnace, and the grate, the efficiency based on coal must necessarily 
be used. 



APPENDIX B 841 

It has been proposed that in reporting the efficiency of a boiler when 
the fuel used contains hydrogen, the efficiency should be considered to 
be the sum of the percentage of the heating value of the fuel which is 
utilized by the boiler in making steam and of the percentage of that 
heating value which is lost in the shape of latent heat in the moisture in 
the chimney gases, which moisture is formed by the burning of the 
hydrogen. This latent heat may amount to over three per cent of the 
total heating value of the fuel. The reason assigned for this proposal 
is that, since it is impossible for this heat to be utilized by the boiler 
because the gases are discharged at a temperature above 212 degrees F. 
it should not be charged against the boiler. The writer does not con- 
sider it advisable that this method of reporting the efficiency should 
be adopted (1) because it is opposed to the generally accepted definition 
of efficiency, which is the useful work received from an apparatus 
divided by the work (or heat value of the fuel) put into it; (2) because 
in order to calculate it it is necessary to know both the percentage of 
hydrogen in the coal and whether or not all of this hydrogen has been 
burned to H 2 0, the first requiring an analysis of the coal, which is not 
always obtainable, and the second an analysis of the gases for hydrogen, 
which cannot be obtained with any approximation to accuracy with our 
present methods of sampling and analyzing gases; and (3) because it 
is opposed to the almost universal custom in reporting boiler tests. It 
is true that the latent heat of the H 2 in the chimney gases cannot be 
utilized (unless an economizer which discharges its gases below 212 
degrees is used), and it is not the fault of the boiler that it cannot be 
utilized. It may be considered the misfortune of the boiler, when tested 
with hydrogenous coal, similar to the misfortune under which an engine 
labors when it is tested while supplied with a condenser which gives a 
vacuum of less than 30 inches of mercury. The engine might give a 
higher efficiency with a vacuum of 30 inches than it would with one of 
27 or 28 inches; but it is not customary to credit the engine with the 
efficiency which it loses on account of the imperfect vacuum. 

Since it is well understood that a boiler cannot show quite as high an 
efficiency (as commonly defined) when using bituminous coal high in 
hydrogen as when using anthracite nearly free from hydrogen, no harm 
is done, and much confusion is avoided, by reporting the efficiency as 
the percentage of the heating value of the coal which is actually utilized 
in making steam. The fact that bituminous coal is used is always 
stated in the report of a test made with that coal. If desired a state- 
ment may also be made in the " heat balance" of the approximate or esti- 
mated percentage of heat which is lost in the latent heat of the moisture 
in the chimney gases, together with the loss due to moisture in the coal. 



842 STEAM POWER PLANT ENGINEERING 

APPENDIX XXV. 

Discrepancy between Commercial and Experimental Results. 

The final result sought by manufacturers, in initiating tests of steam 
or other machinery in actual use, is the value of the work done measured 
in dollars and cents. In some cases the broad question is raised as to 
the saving that may be accomplished by installing improved boilers, 
engines, or other machinery; but more generally it is desired to ascer- 
tain what can be done to produce saving with the apparatus already in 
place under the actual conditions that prevail at the particular location. 
In both these cases it is necessary to ascertain the average cost of the 
work done commercially previous to the test. Frequently, in fact 
generally, this important fact will not be ascertained by an elaborate 
trial, for the reason that everything will be put in order for the test, 
and all details of the trial be conducted so carefully that the losses due 
to average carelessness or want of skill in the past will be eliminated, 
the engineer making the test will not receive proper credit, and the 
owners on seeing the report may conclude that they are already doing 
very well, and perhaps continue old methods with fancied security. If 
the cost of the output of the factory for a given time were ascertained 
in terms of the coal burned during the same time, and compared with 
the corresponding cost for the time of the trial, the latter would fre- 
quently be found to be one-eighth to one-third less than the former, and 
it might not be possible to tell what had caused the difference; for 
instance, whether it was due to putting in order the machinery prior to 
the tests, to greater care exercised by the fireman under the spur of 
careful watching, or whether, as is usually claimed, the coal was different, 
etc., etc. The losses are generally due in the main to the carelessness of 
the firemen. It follows, therefore, that the cost of the power under 
average conditions must be obtained in some quiet way preliminarily. 
Frequently the comparison of the output of the factory with the coal 
burned will not be sufficiently accurate, and it will be necessary to 
devise some corresponding check which will not interfere with the 
regular routine of the establishment. The work of the boilers may be 
checked by arranging a meter so as to continuously measure the feed 
water; and its record, compared with the total weight of coal purchased, 
will frequently give the check desired. Such a check becomes more 
difficult when it is desirable to ascertain the performances of particular 
boilers, and the coal supply is common to all boilers; but by assigning 
particular weighed car loads of coal to the particular boilers, without 
any intimation to the firemen that they are being watched, it may be 



APPENDIX B S43 

possible to ascertain the average performance of the boilers used for 
the particular purpose. Preliminary experiments of this kind conducted 
without notice to employees, and continued through a long period, will 
furnish a basis for comparison with elaborate tests, and it will then be 
possible to point out clearly where the several losses have taken place, 
and the testing engineer will get the credit for the saving shown. 

C. E. E. 

APPENDIX XXXV. 

The Ringelmann Smoke Chart. 

Professor Ringelmann, of Paris, has invented a system of determining 
the relative density or blackness of smoke, which has been communi- 
cated to the writer by Mr. Bryan Donkin, of London, and published in 
Engineering News of November 11, 1897. In making observations of 
the smoke proceeding from a chimney four cards ruled like those in the 
cut, together with a card printed in solid black and another left entirely 
white, are placed in a horizontal row and hung at a point about 50 feet 
from the observer and as nearly as convenient in line with the chimney. 
At this distance the lines become invisible, and the cards appear to be 
of different shades of gray, ranging from very light gray to almost black. 
The observer glances from the smoke coming from the chimney to the 
cards, which are numbered from to 5, determines which card most 
nearly corresponds with the color of the smoke and makes a record 
accordingly, noting the time. Observations should be made continu- 
ously during say one minute, and the estimated average density during 
that minute recorded, and so on, records being made once every minute. 
The average of all the records made during a boiler test is taken as the 
average figure for the smoke density during the test, and the whole of 
the record is plotted on cross-section paper in order to -show how the 
smoke varied in density from time to time. A rule by which the cards 
may be reproduced is given by Professor Ringelmann as follows: 

Card — All white. 

Card 1 — Black lines 1 mm. thick, 10 mm. apart, leaving spaces 9 mm. 
square. 

Card 2 — Lines 2.3 mm. thick, spaces 7.7 mm. square. 
Card 3 — - Lines 3.7 mm. thick, spaces 6.3 mm. square. 
Card 4 — Lines 5.5 mm. thick, spaces 4.5 mm. square. 
Card 5 — All black. 

The cards as printed in Fig. 452 are much smaller than those used by 
Professor Ringelmann, but the thickness and the spacing of the lines 
are the same. w. "K" 



844 STEAM POWER PLANT ENGINEERING 











































































































































































































































































































No.l 

























































































































































































































































































































































































































































































































































































































No.2 






No.3 No.4 

Fig. 452. Ringelman Smoke Chart. 



APPENDIX B 



845 



APPENDIX XXXVIII. 

Chart Showing Graphically the Log of a Trial. 

The well-known method of plotting observations and data on cross- 
section paper and making a chart applying to the test is a useful 
means of representing the exact uniformity of conditions existing 



Lbs CHART SHOWING LOG OF BOILER TEST. 


'■"• 0CO ^ "^ 70 LOS. 


1 1 1 A r-\ l/!\ j y ' 


/"' \ eo 


/\ ^St r omOau 8 «/\ A / \ A / \" 


Z ^=o ^N v^lC^^V^ 


Z l. ZIv Z L3 ^ i_ ^^ 




l±2w r ^ tz^s £ 


m °°° / 0EG . V | ° 


- 14 2 \W* V 




■t r- ^ 1 ^ 


140.000 , 2 -T_ a I 7± 


-_; Zflt _«^s ^ 3 L^ 


.a,^ S ~ 7 2i2-% ^%Xd£^ V 


Z ^ \ IX Ik- ' 


•^ 2 ^ 5 £ ^ 5 ^2 ^ 


Feed Terrtp \ ] .>■' \ 1 \ 




i t f 3 I ] 


mooo ^ |- r--^ a,^<- _j_ 


^ ? 3 / ' ■ » 












„<•" \-^'° 


E • 










te^ ^ - "• oro 


" 80000 - - - - **£ 


i x 


80.000 jr 


^/ 16,000 


£ / 


,/ ^ 


™' m Tfes. 7 


1"' _g . 11.000 


600 ^ •> 




,0,000 _ _ . .. ^ FlueTemp . ' ^ 




60,000 - p's-_~» ^r:_^2._^=2 = =2 = ^U« ! g^_=: = =2 = =:^^: _^^.» 10.000 




5l? f i r_; /.. ' ^'=' . "^-^-°"" C 8000 


«o.ooo -3 T'-s.*^ + ** 1 = **"''° 


^o-^^-o^ 4 ^«g -3.000 


K000 %r ^\ i-\_ +-*~~ 




W.0W ^ r -"j: 2 * N*^ 




/" _. »» "* 


4000 


SO.000 - - ^^^ 








. _< 




— ?:_ _ _ — _ _ 


1 1 M 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 1 o 



Fig. 453, 



during a trial. Such a chart is illustrated in Fig. 453, in which the 
abscissae represent times and the ordinates, on appropriate scales, 
the various observations and data. g. h. b. 






APPENDIX C. 

RULES FOR CONDUCTING STEAM-ENGINE TESTS * 
Code of 1902. 

I. Object of Test. — Ascertain at the outset the specific object of 
the test, whether it be to determine the fulfillment of a contract guar- 
antee, to ascertain the highest economy obtainable, to find the working 
economy and defects under conditions as they exist, to ascertain the 
performance under special conditions, to determine the effect of 
changes in the conditions, or to find the performance of the entire 
boiler and engine plant, and prepare for the test accordingly. 

II. General Condition of the Plant. — Examine the engine and the 
entire plant concerned in the test; note its general condition and any 
points of design, construction, or operation which bear on the objects 
in view. Make a special examination of the valves and pistons for 
leakage by applying the working pressures with the engine at rest, 
and observe the quantity of steam, if any, blowing through per hour. 

If the trial has for an object the determination of the highest 
efficiency obtainable, the valves and pistons must first be made tight, 
and all parts of the engine and its auxiliaries, and all other parts of the 
plant concerned, should be put in the best possible working condition. 

III. Dimensions, etc. — Measure or check the dimensions of the 
cylinders in any case, this being done when they are hot. If they are 
much worn, the average diameter should be determined. Measure 
also the clearance, which should be done, if possible, by filling the 
spaces with water previously measured, the piston being placed at the 
end of the stroke. If the clearance cannot be measured directly, it 
can be determined approximately from the working drawings of the 
cylinder. 

Measure also the dimensions of auxiliaries and accessories, also those 
of the boilers so far as concerned in attaining the objects. It is well 
to supplement these determinations with a sketch or sketches showing 
the general features and arrangement of the different parts of the plant. 

* From the final report of the committee appointed to standardize a system of 
testing steam engines. Trans. A.S.M.E., Vol. XXIV. Greatly abridged. 

846 



APPENDIX C 847 

IV. Coal. — When the trial involves the complete plant, embracing 
boilers as well as engine, determine the character of coal to be used. 
The class, name of the mine, size, moisture, and quality of the coal 
should be stated in the report. It is desirable, for purposes of com- 
parison, that the coal should be of some recognized standard quality 
for the locality where the plant is situated. 

V. Calibration of Instruments. — All instruments and apparatus 
should be calibrated and their reliability and accuracy verified by 
comparison with recognized standards. Such apparatus as is liable to 
change or become broken during a test, as gauges, indicator springs, 
and thermometers, should be calibrated before and after the test. 
The accuracy of scales should be verified by standard weights. When 
a water meter is used, special attention should be given to its cali- 
bration, verifying it both before and after the trial, and, if possible, 
during its progress, the conditions in regard to water pressure and 
rate of flow being made the same in the calibrations as exist through- 
out the trial. 

VI. Leakages of Steam, Water, etc. — In all tests except those of a 
complete plant made under conditions as they exist, the boiler and its 
connections, both steam and feed, as also the steam piping leading to 
the engine and its connections, should, so far as possible, be made 
tight. If absolute tightness cannot be obtained (in point of fact it 
rarely can be), proper allowance should be made for such leakage in 
determining the steam actually consumed by the engine. This, how- 
ever, is not required where a surface condenser is used and the water 
consumption is determined by measuring the discharge of the air 
pump. In such cases it is necessary to make sure that the condenser 
is tight, both before and after the test, against the entrance of circu- 
lating water, or if such occurs to make proper correction for it, deter- 
mining it under the working difference of pressure. Should there 
be excessive leakage of the condenser it should be remedied before 
the test is made. When the steam consumption is determined 
by measuring the discharge of the air pump, any leakage about 
the valve or piston rods of the engine should be carefully guarded 
against. 

Make sure that there is no leakage at any of the connections with 
the apparatus provided for measuring and supplying the feed water 
which could affect the results. All connections should, so far as pos- 
sible, be visible and be blanked off, and where this cannot be done, 
satisfactory assurance should be obtained that there is no leakage 
either in or out. 



848 STEAM POWER PLANT ENGINEERING 

VII. Duration of Test. — The duration cf a test should depend 
largely upon its character and the objects in view. The standard heat 
test of an engine, and, likewise, a test for the simple determination of 
the feed- water consumption, should be continued for at least five 
hours, unless the class of service precludes a continuous run of so long 
duration. It is desirable to prolong the test the number of hours 
stated to obtain a number of consecutive hourly records as a guide in 
analyzing the reliability of the whole. 

Where the water discharged from the surface condenser is measured 
for successive short intervals of time, and the rate is found to be uni- 
form, the test may be of a much shorter duration than where the feed 
water is measured to the boiler. The longer the test with a given set 
of conditions the more accurate the work, and no test should be so 
short that it cannot be divided into several intervals which will give 
results agreeing substantially with each other. 

The commercial test of a complete plant, embracing boilers as well 
as engine, should continue at least one full day of twenty-four hours, 
whether the engine is in motion during the entire time or not. A 
continuous coal test of a boiler and engine should be of at least ten 
hours' duration, or the nearest multiple of the interval between times 
of cleaning fires. 

VIII. Starting and Stopping a Test. — (a) Standard Heat Test and 
Feed-Water Test of Engine : The engine having been brought to the 
normal condition of running, and operated a sufficient length of time 
to be thoroughly heated in all its parts, and the measuring apparatus 
having been adjusted and set to work, the height of water in the 
gauge glasses of the boilers is observed, the depth of water in the 
reservoir from which the feed water is supplied is noted, the exact 
time of day is observed, and the test held to commence. Thereafter 
the measurements determined upon for the test are begun and carried 
forward until its close. If practicable, the test may be commenced at 
some even hour or minute, but it is of the first importance to begin at 
such time as reliable observations of the water heights are obtained, 
whatever the exact time happens to be when these are satisfactorily 
determined. When the time for the close of the test arrives, the 
water should, if possible, be brought to the same height in the glasses 
and to the same depth in the feed- water reservoir as at the beginning, 
delaying the conclusion of the test if necessary to bring about this 
similarity of conditions. If differences occur, the proper corrections 
must be made. 

(6) Complete Engine and Boiler Test : For a continuous running 
test of combined engine or engines, and boiler or boilers, the same 



APPENDIX C 849 

directions apply for beginning and ending the feed-water measurements 
as those just referred to under Section (a). The time of beginning and 
ending such a test should be the regular time of cleaning the fires, and 
the exact time of beginning and ending should be the time when the 
fires are fully cleaned, just preparatory to putting on fresh coal. In cases 
where there are a number of boilers, and it is inconvenient or undesirable 
to clean all fires at once, the time of beginning the test should be deferred 
until they are all cleaned and in a satisfactory state, all the fires being 
then burned down to a uniformly thin condition, the thickness and 
condition being estimated and the test begun just before firing the new 
coal previously weighed. The ending of the test is likewise deferred 
until the fires are all satisfactorily cleaned, being again burned down to 
the same uniformly thin condition as before, and the time of closing 
being taken just before replenishing the fires with new coal. 

For a commercial test of a combined engine and boiler, whether the 
engine runs continuously for the full twenty-four hours of the day or 
only a portion of the time, the fires in the boilers being banked during 
the time when the engine is not in motion, the beginning and ending of 
the test should occur at the regular time of cleaning the fires, the method 
followed being that already given. In cases where the engine is not in 
continuous motion, as, for example, in textile mills, where the working 
time is ten or eleven hours out of the twenty-four, and the fires are 
cleaned and banked at the close of the day's work, the best time for 
starting and stopping a test is the time just before banking, when the 
fires are well burned down and the thickness and condition can be most 
satisfactorily judged. In these, as in all other cases noted, the test 
should be begun by observing the exact time, the thickness and condi- 
tion of the fires on the grates, the height of water in the gauge glasses of 
the boilers, the depth of the water in the reservoir from which the feed 
water is supplied, and other conditions relating to the trial, the same 
observations being again taken at the end of the test, and the conditions 
in all respects being made as nearly as possible the same as at the 
beginning. 

IX. Measurement of Heat Units consumed by the Engine. — The 
measurement of the heat consumption requires the measurement of 
each supply of feed water to the boiler — that is, the water supplied by 
the main feed pump, that supplied by auxiliary pumps, such as jacket 
water, water from separators, drips, etc., and water supplied by gravity 
or other means; also the determination of the temperature of the water 
supplied from each source, together with the pressure and quality of 
the steam. 

The temperatures at the various points should be those applying to 



850 STEAM POWER PLANT ENGINEERING 

the working conditions. The temperature of the feed water should be 
taken near the boiler. This causes the engine to suffer a disadvantage 
from the heat lost by radiation from the pipes which carry the water 
to the boiler, but it is, nevertheless, advisable on the score of simplicity. 
Such pipes would, therefore, be considered a portion of the engine plant. 
This conforms with the rule already recommended for the tests of 
pumping engines where the duty per million heat units is computed 
from the temperature of the feed water taken near the boiler. It 
frequently happens that the measurement of the water requires a change 
in the usual temperature of supply. For example, where the main 
supply is ordinarily drawn from a hot well in which the temperature is 
say 100 degrees F., it may be necessary, owing to the low level of 
the well, to take the supply from some source under a pressure or head 
sufficient to fill the weighing tanks used, and this supply may have a 
temperature much below that of the hot well; possibly as low as 40 
degrees F. The temperature to be used is not the temperature of 
the water as weighed in this case, but that of the working temperature 
of the hot well. The working temperature in cases like this must be 
determined by a special test, and included in the log sheets. 

The heat to be determined is that used by the entire engine equip- 
ment, embracing the main cylinders and all auxiliary cylinders and 
mechanism concerned in the operation of the engine, including the air 
pump, circulating pump, and feed pumps, also the jacket and reheater 
when these are used. No deduction is to be made for steam used by 
auxiliaries unless these are shown by test to be unduly wasteful. In 
this matter an exception should be made in cases of guarantee tests 
where the engine contractor furnishes all the auxiliaries referred to. 
He should, in that case, be responsible for the whole, and no allowance 
should be made for inferior economy, if such exists. Should a deduction 
be made on account of the auxiliaries being unduly wasteful, the method 
of waste and its extent, as compared with the wastes of the main engine 
or other standard of known value, shall be reported definitely. 

The steam pressure and the quality of the steam are to be taken at 
some point conveniently near the throttle, valve. The quantity of 
steam used by the calorimeter must be determined and properly allowed 
for. (See Article XVI, on " Quality of Steam.") 

X. Measurement of Feed Water or Steam Consumption of Engine, 
etc. — The method of determining the steam consumption applicable to 
all plants is to measure all the feed water supplied to the boilers, and 
deduct therefrom the water discharged by separators and drips, as also 
the water and steam which escape on account of leakage of the boiler 
and its pipe connections and leakage of the steam main and branches 



APPENDIX C 851 

connecting the boiler and the engine. In plants where the engine 
exhausts into a surface condenser the steam consumption can be 
measured by determining the quantity of water discharged by the air 
pump, corrected for any leakage of the condenser, and adding thereto 
the steam used by jackets, reheaters, and auxiliaries as determined 
independently. If the leakage of the condenser is too large to satis- 
factorily allow for it, the condenser should, of course, be repaired and 
the leakage again determined before making the test. 

In measuring the water it is best to carry it through a tank or tanks 
resting on platform weighing scales suitably arranged for the purpose, 
the water being afterwards emptied into a reservoir beneath, from which 
the pump is supplied. 

Where extremely large quantities of water must be measured, or in 
some places relatively small quantities, the orifice method of measuring 
is one that can be applied with satisfactory results. In this case the 
average head of water on the orifice must be determined, and, further- 
more, it is important that means should be at hand for calibrating the 
discharge of the orifice under the conditions of use. 

The corrections or deductions to be made for leakage above referred 
to should be applied only to the standard heat-unit test and tests for 
determining simply the steam or feed-water consumption, and not to 
coal tests of combined engine and boiler equipment. In the latter, no 
correction should be made except for leakage of valves connecting to 
other engines and boilers, or for steam used for purposes other than the 
operation of the plant under test. Losses of heat due to imperfections 
of the plant should be charged to the plant, and only such losses as are 
concerned in the working of the engine alone should be charged to the 
engine. 

In measuring jacket water or any supply under pressure which has 
a temperature exceeding 212 degrees F., the water should first be 
cooled, as may be done by discharging it into a tank of cold water 
previously weighed, or by passing it through a coil of pipe submerged 
in running and colder water, preventing thereby the loss of evapo- 
ration which occurs when such hot water is discharged into the 
open air. 

XI. Measurement of Steam used by Auxiliaries. — Although the 
steam used by the auxiliaries — embracing the air pump, circulating 
pump, feed pump, and any other apparatus of this nature, supposing 
them to be steam-driven, also the steam jackets, reheaters, etc., which 
consume steam required for the operation of the engine — is all 
included in the measurement of the steam consumption, as pointed 
out in Article X, yet it is highly desirable that the quantity of steam 



852 STEAM POWER PLANT ENGINEERING 

used by the auxiliaries, and in many cases that used by each auxiliary, 
should be determined exactly, so that the net consumption of the 
main engine cylinders may be ascertained and a complete analysis 
made of the entire work of the engine plant. Where the auxiliary 
cylinders are non-condensing, the steam consumption can often be 
measured by carrying the exhaust for the purpose into a tank of cold 
water resting on scales or through a coil of pipe surrounded by cold 
running water. Another method is to run the auxiliaries as a whole, 
or one by one, from a spare boiler (preferably a small vertical one), 
and measure the feed water supplied to this boiler. The steam used 
by the air and circulating pumps may be measured by running them 
under, as near as possible, the working conditions and speed, the main 
engine and other auxiliaries being stopped, and testing the con- 
sumption by the measuring apparatus used on the main trial. For a 
short trial, to obtain approximate results, measurement can be made 
by the water-gauge glass method, the feed supply being shut off. 
When the engine has a surface condenser, the quantity of steam used 
by the auxiliaries may be ascertained by allowing the engine alone to 
exhaust into the condenser, measuring the feed water supplied to the 
boiler and the water discharged by the air pump, and subtracting one 
from the other, after allowing for losses by leakage. 

XII. Coal Measurement. — (a) Commercial Tests : In commercial 
tests of the combined engine and boiler equipment, or those made 
under ordinary conditions of commercial service, the test should, as 
pointed out in Article VII, extend over the entire period of the day; 
that is, twenty-four hours, or a number of days of that duration. 
Consequently, the coal consumption should be determined for the 
entire time. If the engine runs but a part of the time, and during the 
remaining portion the fires are banked, the measurement of coal should 
include that used for banking. It is well, however, in such cases, to 
determine separately the amount consumed during the time the 
engine is in operation and that consumed during the period while the 
fires are banked, so as to have complete data for purposes of analysis 
and comparison, using suitable precautions to obtain reliable measure- 
ments. The measurement of coal begins with the first firing, after 
cleaning the furnaces and burning down at the beginning of the test, 
as pointed out in Article VIII, and ends with the last firing, at the 
expiration of the allotted time. 

(b) Continuous Running Tests : In continuous running tests which, 
as pointed out in Article VII, cover one or more periods which elapse 
between the cleaning of the fires, the same principle applies as that 
mentioned under the above heading (a); viz., the coal measurement 



APPENDIX C 853 

begins with the first firing, after cleaning and burning down, and the 
measurement ends with the last firing, before cleaning and burning 
down at the close of the trial. 

(c) Coal Tests in General : When not otherwise specially under- 
stood, a coal test of a combined engine and boiler plant is held to refer 
to the commercial test above noted, and the measurement of coal 
should conform thereto. 

In connection with coal measurements, whatever the class of tests, 
it is important to ascertain the percentage of moisture in the coal, 
the weight of ashes and refuse, and, where possible, the approximate 
and ultimate analysis of the coal, following all the methods and details 
advocated in the latest report of the Boiler Test Committee of the 
Society. (See Vol. XXI, p. 34.) 

(d) Other Fuels than Coal : For all other solid fuels than coal the 
same directions in regard to measurement should be followed as those 
given for coal. If the boilers are run with oil or gas, the measure- 
ments relating to stopping and starting are much simplified, because 
the fuel is burned as fast as supplied, and there is no body of fuel con- 
stantly in the furnace, as in the case of using solid fuel. When oil 
is used, it should be weighed, and when gas is used, it should be 
measured in a calibrated gas meter or a gasometer. 

XIII. Indicated Horse Power. — The indicated horse power should 
be determined from the average mean effective pressure of diagrams 
taken at intervals of twenty minutes, and at more frequent intervals 
if the nature of the test makes this necessary, for each end of each 
cylinder. With variable loads, such as those of engines driving gener- 
ators for electric railroad work, and of rubber-grinding and rolling-mill 
engines, the diagrams cannot be taken too often. In cases like the 
latter, one method of obtaining suitable averages is to take a series of 
diagrams on the same blank card without unhooking the driving cord, 
and apply the pencil at successive intervals of ten seconds until two 
minutes' time or more has elapsed, thereby obtaining a dozen or more 
indications in the time covered. This tends to insure the determina- 
tion of a fair average for that period. In taking diagrams for variable 
loads, as indeed for any load, the pencil should be applied long enough 
to cover several successive revolutions, so that the variations produced 
by the action of the governor may be properly recorded. To 
determine whether the governor is subject to what is called " racing " 
or " hunting," a " variation diagram " should be obtained; that is, 
one in which the pencil is applied a sufficient time to cover a complete 
cycle of variations. When the governor is found to be working in this 
manner, the defect should be remedied before proceeding with the test. 



854 STEAM POWER PLANT ENGINEERING 

It is seldom necessary, as far as average power measurements are 
concerned, to obtain diagrams at precisely the same instant at the 
two ends of the cylinder, or at the same instant on all the cylinders, 
when there are more than one. All that is required is to take the 
diagrams at regular intervals. Should the diagrams vary so much 
among themselves that the average may not be a fair one, it signifies 
that they should be taken more frequently, and not that special care 
should be employed to obtain the diagrams of each set at precisely 
"the same time. When diagrams are taken during the time when the 
engine is working up to speed at the start, or when a study of valve 
setting and steam distribution is being made, they should be taken at 
as nearly the same time as practicable. In cases where the diagrams 
are to be taken simultaneously, the best plan is to have an operator 
stationed at each indicator. This is desirable, even where an electric 
or other device is employed to operate all the instruments at once ; 
for unless there are enough operators, it is necessary to open the indi- 
cator cocks some time before taking the diagrams and run the risk of 
clogging the pistons and heating the high-pressure springs above the 
ordinary working temperature. 

The most satisfactory driving rig for indicating seems to be some 
form of well-made pantagraph, with driving cord of fine annealed wire 
leading to the indicator. The reducing motion, whatever it may be, 
and the connections to the indicator, should be so perfect as to pro- 
duce diagrams of equal lengths when the same indicator is attached to 
either end of the cylinder, and produce a proportionate reduction of the 
motion of the piston at every point of the stroke, as proved by test. 

The use of a three-way cock and a single indicator connected to the 
two ends of the cylinder is not advised, except in cases where it is 
impracticable to use an indicator close to each end. If a three-way 
cock is used, the error produced should be determined and allowed for. 

To determine the average power developed in cases where the engine 
starts from rest during the progress of the trial, as in a commercial 
test of a plant where the engine runs only a portion of the twenty- 
four hours, a number of diagrams should be taken during the period of 
getting up speed and applying the working load, the corresponding 
speed for each set of diagrams being counted. The power shown by 
these diagrams for the proportionate time should be included in the 
average for the whole run, and the duration should be the time the 
throttle valve is open. 

XIV. Testing Indicator Springs. — To make a perfectly satisfactory 
comparison of indicator springs with standards, the calibration should 
be made, if this were practical, under the same conditions as those 



APPENDIX C 855 

pertaining to their ordinary use. Owing to the fact that the pressure 
of the steam in the indicator cylinder and the corresponding temperature 
are undergoing continual changes, it becomes almost impossible to 
compare the springs with any standard under such conditions. There 
must be a constant pressure during the time that the comparison is 
being made. Although the best that can be done is not altogether 
satisfactory, it seems that we must be content with it. To bring the 
conditions as nearly as possible to those of the working indicator, the 
steam should be admitted to the indicator as short a time as practicable 
for each of the pressures tried, and then the indicator cock should be 
closed and the steam exhausted therefrom before another pressure is 
tried. By this means the parts are heated and cooled somewhat the 
same as under the working conditions. We recommend, therefore, that 
for each required pressure the first step be to open and close the indicator 
cock a number of times in quick succession, then to quickly draw the 
line on the paper for the desired record, observing the gauge or other 
standard at the instant when the line is drawn. A corresponding 
atmospheric line is taken immediately after obtaining the line at the 
given pressure, so as to eliminate any difference in the temperature of 
the parts of the indicator. This appears to be a better method (although 
less readily carried on and requiring more care) than the one heretofore 
more commonly used, where the indicator cock is kept continually open 
and the pressure is gradually rising or falling through the range of 
comparison. 

The calibration .should be made for at least five points, two of these 
being for the pressure corresponding as near as may be to the initial 
and back pressures, and three for intermediate points equally distant. 

For pressures above the atmosphere, the proper standard recom- 
mended is the dead-weight testing apparatus, or a reliable mercury 
column, or an accurate steam gauge proved correct, or of known error, 
by either of these standards. For pressures below the atmosphere the 
best standard to use is a mercury column. 

The correct scale of spring to be used for working out the mean 
effective pressure of the diagrams should be the average based on the 
calibration, and this may be ascertained in the manner pointed out 
below. 

XV. Brake Horse Power. — This term applies to the power delivered 
from the fly-wheel shaft of the engine. It is the power absorbed by a 
friction brake applied to the rim of the wheel or to the shaft. A form 
of brake is preferred that is self-adjusting to a certain extent, so that it 
will of itself tend to maintain a constant resistance at the rim of the 
wheel. One of the simplest brakes for comparatively small engines, 



856 



STEAM POWER PLANT ENGINEERING 



which may be made to embody this principle, consists of a cotton or 
hemp rope, or a number of ropes, encircling the wheel, arranged with 
weighing scales or other means for showing the strain. An ordinary 
band brake may also be constructed so as to embody the principle. 
The wheel should be provided with interior flanges for holding water 
used for keeping the rim cool. 




-^^^^^^^^^^^^^^^^^^^^^^^^^^^^^^j^^^^^^^^- 



Fig. 454. 



Rope Brakes. 



Fig. 455. 



A self-adjusting rope brake is illustrated in Fig. 454, where it will 
be seen that, if the friction at the rim of the wheel increases, it will 
lift the weight A, which action will diminish the tension in the end B 
of the rope and thus prevent a further increase in the friction. The 
same device can be used for a band brake of the ordinary construction. 
Where space below the wheel is limited, a cross bar, C, supported by 
a chain tackle exactly at its center point may be used as shown in Fig. 
455, thereby causing the action of the weight on the brake to be up- 
ward. A safety stop should be used with either form, to prevent the 
weights being accidentally raised more than a certain amount. 

The water-friction brake is specially adapted for high speeds and has 
the advantage of being self-cooling. The Alden brake is also self- 
cooling and is capable of fine adjustment. 

A water-friction brake is shown in Fig. 456. It consists of two cir- 
cular disks, A and B, attached to the shaft C, and revolving in a case, 
E, between fixed planes. The space between the disks and planes is 






APPENDIX C 



857 



supplied with running water, which enters at D and escapes at the 
cocks F, G, and H. The friction of the water against the surfaces con- 
stitutes a resistance which absorbs the desired power, and the heat 
generated within is carried away by the water itself. The water is 
thrown outward by centrifugal action and fills the outer portion of 
the case. The greater the depth of the ring of water the greater the 




Fig. 456. Alden Absorption Dynamometer. 

amount of power absorbed. By suitably adjusting the amount of 
water entering and leaving any desired power can be obtained. 
Water-friction brakes have been used successfully at speeds of over 
20,000 revolutions per minute. 

For description of the Alden brake, see Transactions, Vol. XI, p. 958. 

XVI. Quality of Steam. — When ordinary saturated steam is used, 
its quality should be obtained by the use of a throttling calorimeter 
attached to the main steam pipe near the throttle valve. When the 
steam is superheated, the amount of superheating should be found by 
the use of a thermometer placed in a thermometer-well filled with mer- 
cury, inserted in the pipe. The sampling pipe for the calorimeter should, 
if possible, be attached to a section of the main pipe having a vertical 
direction, with the steam preferably passing upward, and the sampling 
nozzle should be made of a half -inch pipe, having at least 20 one-eighth- 
inch holes in its perforated surface. The readings of the calorimeter 
should be corrected for radiation of the instrument, or they should be 
referred to a normal reading, as pointed out below. If the steam is 



858 STEAM POWER PLANT ENGINEERING 

superheated, the amount of superheating should be obtained by refer- 
ring the reading of the thermometer to that of the same thermometer 
when the steam within the pipe is saturated, and not by taking the 
difference between the reading of the thermometer and the temper- 
ature of saturated steam at the observed pressure as given in a steam 
table. 

XVII. Speed. — There are several reliable methods of ascertaining 
speed, or the number of revolutions of the engine crank-shaft per 
minute. The simplest is the familiar method of counting the number 
of turns for a period of one minute with the eye fixed on the second 
hand of a timepiece. Another is the use of a counter held for a 
minute or a number of minutes against the end of the main shaft. 
Another is the use of a reliable calibrated tachometer held likewise 
against the end of the shaft. The most reliable method, and the one 
we recommend, is the use of a continuous recording engine register or 
counter, taking the total reading each time that the general test data 
are recorded, and computing the revolutions per minute corresponding 
to the difference in the readings of the instrument. When the speed 
is above 250 revolutions per minute, it is almost impossible to make a 
satisfactory counting of the revolutions without the use of some form 
of mechanical counter. 

The determination of variation of speed during a single revolution, 
or the effect of the fluctuation due to sudden changes of the load, is 
also desirable, especially in engines driving electric generators used for 
lighting purposes. There is at present no recognized standard method 
of making such determinations, and if such are desired, the method 
employed may be devised by the person making the test and described 
in detail in the report. 

XVIII. Recording the Data. — Take note of every event connected 
with the progress of the trial whether it seems at the time to be 
important or unimportant. Record the time of every event and 
time of taking every weight and every observation. Observe the 
pressures, temperatures, water heights, speeds, etc., every twenty or 
thirty minutes when the conditions are practically uniform, and at 
much more frequent intervals if the conditions vary. Observations 
which concern the feed-water measurements should be made with 
special care at the expiration of each hour of the trial, so as to divide 
the tests into hourly periods and show the uniformity of the conditions 
and results as the test goes forward. Where the water discharged from 
a surface condenser is weighed it may be advisable to divide the test 
by this means into periods of less than one hour. 



. APPENDIX C 859 

The data and observations of the test should be kept on properly 
prepared blanks or in notebooks containing columns suitably arranged 
for a clear record. As different observers have their own individual 
ideas as to how such records should be kept, no special form of log 
sheet is given as a necessary part of the code. 

XIX. Uniformity of Conditions. — In a test having for an object 
the determination of the maximum economy obtainable from an 
engine, or where it is desired to ascertain with special accuracy the 
effect of predetermined conditions of operation, it is important that all 
the conditions under which the engine is operated should be main- 
tained uniformly constant. This requirement applies especially to the 
pressure, the speed, the load, the rate of feeding the various supplies of 
water, the height of water in the gauge glasses, and the depth of water 
in the feed-water reservoir. 

XX. Analysis of Indicator Diagrams. — (a) Steam accounted for by 
the Indicator : The simplest method of computing the steam accounted 
for by the indicator is the use of the formula 

M = 4 3 I 5 ^ [«? + E ) X Wc-(H + E) X Wh], 

which gives the weight in pounds per indicated horse power per hour. 
In this formula the symbol " M.E.P." refers to the mean effective 
pressure. In multiple-expansion engines this is the combined mean 
effective pressure referred to the cylinder in question. The symbol C 
refers to the proportion of the stroke completed at points on the 
expansion line of the diagram near the actual cut-off or release; the 
symbol H to the proportion of compression; and the symbol E to 
the proportion of clearance; all of which are determined from the indi- 
cator diagram. The symbol Wc refers to the weight of one cubic foot 
of steam at the cut-off or release pressure; and the symbol Wh to the 
weight of one cubic foot of steam at the compression pressure; these 
weights being taken from steam tables of recognized accuracy. The 
points near the cut-off and release on the expansion line and the point 
on the compression line are located as shown on the sample diagram, 
Fig. 457. They are the points in the case of the expansion and com- 
pression lines of the diagram which mark the complete closure of the 
valve. The point near the cut-off, for example, lies where the curve 
of expansion begins after the rounding of the diagram due to the wire- 
drawing which occurs while the valve is closing. This cut-off may be 
located by finding the point where the curve is tangent to a hyper- 
bolic curve. 



860 STEAM POWER PLANT ENGINEERING 

Should the point in the compression curve be at the same height as 
the point in the expansion curve, then Wc = Wh, and the formula 
becomes 

in which (C — H) represents the distance between the two points 
divided by the length of the diagram. 

When the load and all other conditions are substantially uniform, it 
is unnecessary to work up the steam accounted for by the indicator 




Compression 



Atmospheric Line 

Fig. 457. Showing Points where " Steam Accounted for by Indicator "is 

Computed. 

from all the diagrams taken. Five or more sample diagrams may be 
selected and the computations based on the samples instead of on the 
whole. 

(b) Sample Indicator Diagrams : In order that the report of a test 
may afford complete information regarding the conditions of the test, 
sample indicator diagrams should be selected from those taken and 
copies appended to the tables of results. In cases where the engine is 
of the multiple-expansion type these sample diagrams may also be 
arranged in the form of a " combined " diagram. 

(c) The Point of Cut-off: The term " cut-off" as applied to steam 
engines, although somewhat indefinite, is usually considered to be at an 
earlier point in the stroke than the beginning of the real expansion line. 
That the cut-off point may be defined in exact terms for commercial 
purposes, as used in steam-engine specifications and contracts, the 
Committee recommends that, unless otherwise specified, the commercial 
cut-off, which seems to be an appropriate expression for this term, be 
ascertained as follows: Through a point showing the maximum pressure 



APPENDIX C 



861 



during admission draw a line parallel to the atmospheric line. Through 
the point on the expansion line near the actual cut-off, referred to in 
Section XX (a), draw a hyperbolic curve. The point where these two 
lines intersect is to be considered the commercial cut-off point. The 
percentage is then found by dividing the length of the diagram measured 
to this point by the total length of the diagram and multiplying the 
result by 100. 

E C B A 




Four Valve Engine, Slow Speed, Commercial 



Cut-off= I?. 



The principle involved in locating the commercial cut-off is shown in 
Figs. 458 and 459, the first of which represents a diagram from a slow- 
speed Corliss engine and the second a diagram from a single-valve 
high-speed engine. In the latter case where, owing to the fling of the 




Fig. 459. Single Valve Engine, High-Speed, Com- 
BC 



mercial Cut-off = 



AC" 



pencil, the steam line vibrates, the maximum pressure is found by 
taking a mean of the vibrations at the highest point. 

The commercial cut-off as thus determined is situated at an earlier 



862 STEAM POWER PLANT ENGINEERING 

point of the stroke than the actual cut-off referred to in computing the 
" steam accounted for " by the indicator in Section XX (a). 

(d) Ratio of Expansion : The ratio of expansion for a simple engine 
is determined by dividing the volume corresponding to the piston dis- 
placement, including clearance, by the volume of the steam at the com- 
mercial cut-off, including clearance. 

In a multiple-expansion engine it is determined by dividing the net 
volume of the steam indicated by the low-pressure diagram at the end 
of the expansion line, assumed to be continued to the end of the stroke, 
by the net volume of the steam at the maximum pressure during admis- 
sion to the high-pressure cylinder. 

(e) Diagram Factor : The diagram factor is the proportion borne by 
the actual mean effective pressure measured from the indicator diagram 
to that of a diagram in which the various operations of admission, 
expansion, release, and compression are carried on under assumed con- 
ditions. The factor recommended refers to an ideal diagram which 
represents the maximum power obtainable from the steam accounted 
for by the indicator diagrams at the point of cut-off, assuming first that 
the engine has no clearance; second, that there are no losses through 
wire-drawing the steam either during the admission or the release; 
third, that the expansion line is a hyperbolic curve; and fourth, that the 
initial pressure is that of the boiler and the back pressure that of the 
atmosphere for a non-condensing engine and of the condenser for a 
condensing engine. 

The diagram factor is useful for comparing the steam distribution 
losses in different engines, and is of special use to the engine designer, 
for by multiplying the mean effective pressure obtained from the 
assumed theoretical diagrams by it he will obtain the actual mean 
effective pressure that should be developed in an engine of the type 
considered. The expansion and compression curves are taken as 
hyperbolas, because such curves are ordinarily used by engine builders 
in their work, and a diagram based on such curves will be more useful to 
them than one where the curves are constructed according to a more 
exact law. 

In cases where there is a considerable loss of pressure between the 
boiler and the engine, as where steam is transmitted from a central 
plant to a number of consumers, the pressure of the steam in the supply 
main should be used in place of the boiler pressure in constructing the 
diagrams. 

XXI. Standards of Economy and Efficiency. — The hourly consump- 
tion of heat, determined by employing the actual temperature of the 
feed water to the boiler, as pointed out in Article IX of the Code, divided 



APPENDIX C 863 

by the indicated and brake horse power, that is, the number of heat 
units consumed per indicated and per brake horse power per hour, is 
the standard of engine efficiency recommended by the Committee. 
The consumption per hour is chosen rather than the consumption per 
minute, so as to conform with the designation of time applied to the 
more familiar units of coal and water measurement which have hereto- 
fore been used. The British standard, where the temperature of the 
feed water is taken as that corresponding to the temperature of the 
back-pressure steam, allowance being made for any drips from jackets 
or reheaters, is also included in the tables. 

It is useful in this connection to express the efficiency in its more 
scientific form, or what is called the " thermal efficiency ratio." The 
thermal efficiency ratio is the proportion which the heat equivalent of 
the power developed bears to the total amount of heat actually con- 
sumed, as determined by test. The heat converted into work repre- 
sented by one horse power is 1,980,000 foot-pounds per hour, and this 
divided by 778 equals 2545 British thermal units. Consequently the 
thermal efficiency ratio is expressed by the fraction 

2545 



2545 

B.T.U. per H.P. per hour' 



XXII. Heat Analysis. — For certain scientific investigations it is 
useful to make a heat analysis of the diagram to show the interchange 
of heat from steam to cylinder walls, etc., which is going on within the 
cylinder. This is unnecessary for commercial tests. 

XXIII. Temperature- Entropy Diagram. — The study of the heat 
analysis is facilitated by the use of the temperature-entropy diagram 
in which areas represent quantities of heat, the coordinates being 
the absolute temperature and entropy. Such a diagram is shown in 
Fig. 460. 

When the quantities given in the steam tables are plotted, two 

curves, AA and BE, are obtained which may be termed the water line 

and the steam line, AA being the logarithmic curve if the specific heat 

of the water is taken as constant. The diagram refers to a unit weight 

of the agent, and the heat necessary to raise a pound of water from 

the temperature ma to the temperature pa' and evaporate it at 

that temperature is represented by the area aa'b'qm. If the steam 

be now expanded adiabatically the temperature will fall to qs and 

as 
x per cent = - — will remain as steam, the rest being liquefied. If the 
ab 

steam is now rejected, it carries away with it the heat sqma, the work 



864 



STEAM POWER PLANT ENGINEERING 



area being a'b'sa, from which must be deducted the work w (ex- 
pressed in heat units) to pump a pound of water into the boiler. The 
efficiency of this cycle is evidently 



in which 



X = 



+ L l — xL 2 — w 
h + L t ' 



, T i Li 

log c — + — 
ar + p'y 6c T 2 T. x 

ab L„ 



By the action of the walls a portion of the steam is liquefied prior 
to the expansion, which therefore begins at e, and since the cooling 






Fig. 460. Fig. 461. 

Temperature-Entropy Diagrams. 

action of the walls continues, the expansion line falls off to ef, from 
which point a reverse action takes place and the expansion line bends 






APPENDIX C 865 

over to g. Finally, since the release takes place before the condenser 
temperature is reached, the heat rejection starts at g, following a line 
of equal volume until the exhaust port temperature is reached at /. 
If heat is added during expansion enough to keep the steam theo- 
retically saturated, as, for example, by a water jacket, such additional 
heat is represented by the area b'bnq, and the additional work obtained 
by the triangle b'bs. If the steam is superheated sufficiently to give 
by expansion theoretically dry steam at the end, such additional heat 
is represented by the area b'vnq and the additional work by b'vbs. 
Neither of these extra amounts of work is realized in practice, and 
it is evident from the diagram that the heat thus applied is in both 
cases less efficient than in the principal cycle. Nevertheless the action 
in each case is to bring the point e nearer the point &' and to effect a 
notable net economy. 

The Carnot cycle would be obtained if in the Rankine cycle the 
rejection of heat were stopped at r and the temperature of the mix- 
ture raised to a' by compression. This cannot be practically accom- 
plished, but a system of feed-water heaters has been suggested and 
exemplified in the Nordberg engine, which is theoretically a close 
equivalent to it. Where steam is expanded in say three cylinders, the 
feed water may be successively heated from the receiver intermediate 
between each pair, the effect of which is illustrated in Fig. 461. The 
expansion line follows the heavy line,, being carried over to y by the 
first feed-water heater and to y' by the second feed-water heater. With 
an infinite number of such feed-water heaters, the line yy f would be 
parallel to aa', and the cycle equivalent to that of Carnot. 

XXIV. Ratio of Economy of an Engine to that of an Ideal Engine. — 
The ideal engine recommended for obtaining this ratio is that which 
was adopted by the Committee appointed by the Civil Engineers of 
London to consider and report a standard thermal efficiency for 
steam engines. This engine is one which follows the Rankine cycle, 
where steam at a constant pressure is admitted into the cylinder 
with no clearance, and after the point of cut-off is expanded adiabat- 
ically to the back pressure. In obtaining the economy of this engine 
the feed water is assumed to be returned to the boiler at the exhaust 
temperature. Such a cycle is preferable to the Carnot for the purpose 
at hand, because the Carnot cycle is theoretically impossible for an 
engine using superheated steam produced at a constant pressure, and 
the gain in efficiency for superheated steam corresponding to the 
Carnot efficiency will be much greater than that possible for the actual 
cycle. 

The ratio of the economy of an engine to that of the ideal 
engine is obtained by dividing the heat consumption per indicated 



866 STEAM POWER PLANT ENGINEERING 

horse power per minute for the ideal engine by that of the actual 
engine. 

XXV. Miscellaneous. — In the case of tests of combined engines 
and boiler plants, where the full data of the boiler performance is to 
be determined, reference should be made to the directions given by the 
Boiler Test Committee of the Society, Code of 1899. (See Vol. XXI, 
p. 34.) 

In tests made for scientific research, and in those made on special 
forms of engines, the line of procedure must be varied according to the 
special objects in view, and it has been deemed unnecessary to go into 
particulars applying to such tests. 

In testing steam pumping engines and locomotives in accordance 
with the standard methods of conducting such tests, recommended by 
the committees of the Society, reference should be made to the reports 
of those committees in the Transactions, Vol. XII, p. 530, and in Vol. 
XIV, p. 1312. 

XXVI. Report of Test. — The data and results of the test should be 
reported in the manner and in the order outlined in one of the fol- 
lowing tables, the first of which gives, it is hoped, a complete sum- 
mary of all the data and results as applied not only to the standard 
heat-unit test but also to tests of combined engine and boiler for deter- 
mining all questions of performance, whatever the class of service; 
the second refers to a short form of report giving the necessary data 
and results for the standard heat test; and the third to a short form 
of report for a feed-water test. It is the intention that the tables 
should be full enough to apply to any type of engine, but where not 
so, or where special data and results are determined, additional results 
may be inserted under the appropriate headings. Although these 
forms are arranged so as to be used for expressing the principal data 
and results of tests of pumping engines and locomotives, as well as 
for all other classes of steam engines, it is not the intention that they 
shall supplant the forms recommended by the committees on Duty 
Trials and Locomotives in cases where the full report of a test of such 
engines is desired. _ 

It is recommended that any report be supplemented by a chart in 
which the data of the test are graphically presented. (As an example 
of such a chart as applied to a boiler test, see Vol. XXI, p. 104.) 

TABLE NO. 1. 
Not reprinted here. See Trans. A.S.M.E. 24-702. 



APPENDIX C 867 

TABLE NO. 2. 

DATA AND RESULTS OF STANDARD HEAT TEST OF STEAM ENGINE. 

Arranged according to the Short Form advised by the Engine Test Committee of the 
American Society of Mechanical Engineers. Code of 1902. 

1. Made by of 

on engine located at 

to determine 

2. Date of trial 

3. Type and class of engine ; also of condenser 



4. Dimensions of main engine 

(a) Diameter of cylinder in. 

(b) Stroke of piston ft. 

(c) Diameter of piston rod in. 

(d) Average clearance p. c. 

(e) Ratio of volume of cylinder to high- 

pressure cylinder 

(/) Horse-power constant for one pound mean 
effective pressure and one revolution 
per minute 

5. Dimensions and type of auxiliaries 



1st Cyl. 2d Cyl. 3d Cyl. 



Total Quantities, Time, etc. 

6. Duration of test hours. 

7. Total water fed to boilers from main source of supply pounds. 

8. Total water fed from auxiliary supplies : 

(a) 

(6) 

(0 

9. Total water fed to boilers from all sources pounds. 

10. Moisture in steam or superheating near throttle p. c. or deg. 

11. Factor of correction for quality of steam 

12. Total dry steam consumed for all purposes pounds 

Hourly Quantities. 

13. Water fed from main source of supply " 



14. Water fed from auxiliary supplies: 

(a) 

(b) 

(c) 

15. Total water fed to boilers per hour 

16. Total dry steam consumed per hour 

17. Loss of steam and water per hour due to drips from main steam 

pipes and to leakage of plant 

18. Net dry steam consumed per hour by engine and auxiliaries 



868 



STEAM POWER PLANT ENGINEERING 



Pressures and Temperatures (Corrected). 

19. Pressure in steam pipe near throttle by gauge lb. per sq. in. 

20. Barometric pressure of atmosphere in inches of mercury.. . inches. 

21. Pressure in receivers by gauge lb. per sq. in. 

22. Vacuum in condenser in inches of mercury inches. 

23. Pressure in jackets and reheaters by gauge lb. per sq. in. 

24. Temperature of main supply of feed water degrees F. 

25. Temperature of auxiliary supplies of feed water: 

(«) 

(&) 

(c) 

26. Ideal feed-water temperature corresponding to pressure of steam 

in the exhaust pipe, allowance being made for heat derived 

from jacket or reheater drips " 



Data Relating to Heat Measurement. 

27. Heat units per pound of feed water, main supply 

28. Heat units per pound of feed water, auxiliary supplies : 

(«) 

(b) 

(c) 

29. Heat units consumed per hour, main supply 

30. Heat units consumed per hour, auxiliary supplies: 



B.T.U. 



31. 
32. 
33. 



34. 



(a) 
(6) 
(c). 



Total heat units consumed per hour for all purposes 

Loss of heat per hour due to leakage of plant, drips, etc 

Net heat units consumed per hour: 

(a) By engine alone 

(b) By auxiliaries 

Heat units consumed per hour by engine alone, reckoned from 

temperature given in line 26 



35. 

36. 



38. 



39. 



Indicator Diagrams. 

Commercial cut-off in per cent of stroke 

Initial pressure in pounds per square inch above 

atmosphere 

Back pressure at mid-stroke above or below at- 
mosphere in pounds per square inch 

Mean effective pressure in pounds per square 

inch 

Equivalent mean effective pressure in pounds 
per square inch : 

(a) Referred to first cylinder 

(6) Referred to second cylinder 

(c) Referred to third cylinder 



1st Cyl. 



2d Cyl. 3d Cyl. 



APPENDIX C 869 

IstCyl. 2dCyl. 3d Cyl. 

40. Pressures and percentages used in computing 

the steam accounted for by the indicator 
diagrams, measured to points on the expan- 
sion and compression curves 

Pressure above zero in pounds per square inch: 

(a) Near cut-off 

(6) Near release 

(c) Near beginning of compression 

Percentage of stroke at points where pressures 
are measured: 

(a) Near cut-off 

(6) Near release 

(c) Near beginning of compression 

41. Steam accounted for by indicator in pounds per 

I.H.P. per hour: 

(a) Near cut-off 

(6) Near release 

42. Ratio of expansion 

Speed 

43. Revolutions per minute revolutions. 

Power. 
44:. Indicated horse power developed by main-engine cylinders: 

First cylinder horse power. 

Second cylinder " 

Third cylinder " 

Total 

45. Brake horse power developed by engine " 

Standard Efficiency and other Results * 

46. Heat units consumed by engine and auxiliaries per hour: 

(a) per indicated horse power B.T.TJ. 

(6) per brake horse power " 

47. Equivalent standard coal in pounds per hour: 

(a) per indicated horse power pounds. 

(6) per brake horse power " 

48. Heat units consumed by main engine per hour corresponding to 

ideal maximum temperature of feed water given in line 26, 
British standard: 

(a) per indicated horse power B.T.U. 

(b) per brake horse power " 

49. Dry steam consumed per indicated horse power per hour: 

(a) Main cylinders, including jackets pounds. 

(6) Auxiliary cylinders " 

(c) Engine and auxiliaries " 

* The horse power referred to above (items 46-50) is that of the main engine, 
exclusive of auxiliaries. 



870 STEAM POWER PLANT ENGINEERING 

50. Dry steam consumed per brake horse power per hour: 

(a) Main cylinders, including jackets pounds. 

(6) Auxiliary cylinders " 

(c) Engine and auxiliaries " 

51. Percentage of steam used by main-engine cylinders accounted 

for by indicator diagrams, near cut-off of high-pressure cylinder per cent. 

Additional Data. 

Add any additional data bearing on the particular objects of the test or relating 
to the special class of service for which the engine is used. Also give copies of 
indicator diagrams nearest the mean, and the corresponding scales. 



TABLE NO. 3. 

DATA AND RESULTS OF FEED-WATER TEST OF STEAM ENGINE. 

Arranged according to the Short Form advised by the Engine Test Committee of 
the American Society of Mechanical Engineers. Code of 1902. 



1. Made by of , 

on engine located at 

to determine 



2. Date of trial 

3. Type of engine (simple, compound, or other multiple expansion; condensing 

or non-condensing) 

4. Class of engine (mill, marine, locomotive, pumping, electric, or other) 

5. Rated power of engine 

6. Name of builders 

7. Number and arrangement of cylinders of engine; how lagged; type of valves 

and of condensers 

8. Dimensions of engine lst CyL 2d CyL 3d CyL 

(a) Single or double acting 

(6) Cylinder dimensions: 

Bore in. 

Stroke , ft. 

Diameter of piston rod in. 

Diameter of tail rod in. 

(c) Clearance in per cent of volume displaced 

by piston per stroke: 

Head end 

Crank end 

Average 

(d) Ratio of volume of each cylinder to 

volume of high-pressure cylinder. . . . 

(e) Horse-power constant for one pound mean 

effective pressure and one revolution 
per minute 



APPENDIX C 871 

Total Quantities, Time, etc. 
9. Duration of test hours. 

10. Water fed to boilers from main source of supply pounds. 

11. Water fed from auxiliary supplies: 

(a) 

(b) 

(c) 

12. Total water fed from all sources 

13. Moisture in steam or superheating near throttle * p. c. or deg. 

14. Factor of correction for quality of steam 

15. Total dry steam consumed for all purposes pounds. 

Hourly Quantities. 

16. Water fed from main source of supply " 

17. Water fed from auxiliary supplies: 

(a) 

(6) 

(c) 

18. Total water fed to boilers per hour " 

19. Total dry steam consumed per hour " 

20. Loss of steam and water per hour due to leakage of plant, drips, 

etc ' 

21. Net dry steam consumed per hour by engine and auxiliaries " 

22. Dry steam consumed per hour: 

(a) Main cylinders " 

(b) Jackets and reheaters " 

Pressures and Temperatures (Corrected). 

23. Steam pipe pressure near throttle, by gauge lb. per sq. in. 

24. Barometric pressure of atmosphere in inches of mercury inches. 

25. Pressure in first receiver by gauge lb. per sq. in. 

26. Pressure in second receiver by gauge 

27. Vacuum in condenser: 

(a) In inches of mercury inches. 

(b) Corresponding total pressure lb. per sq. in. 

28. Pressure in steam jackets by gauge lb. per sq. in. 

29. Pressure in reheater by gauge 

30. Superheating of steam in first receiver degrees F. 

31. Superheating of steam in second receiver 

Indicator Diagrams. 

1st Cyl. 2d Cyl. 3d Cyl. 

32. Commercial cut-off in per cent of stroke 

33. Initial pressure in pounds per square inch above 

atmosphere 

* In case of superheated steam engines, determine, if practicable, the temperature 
of the steam in each cvlinder. 



872 STEAM POWER PLANT ENGINEERING 

IstCyl. 2dCyl. 3dCyl. 

34. Back pressure at mid-stroke above or below 

atmosphere in pounds per square inch 

35. Mean effective pressure in pounds per square 

inch 

36. Equivalent mean effective pressure in pounds per 

square inch per indicated horse power 

(a) Referred to first cylinder. 

(b) Referred to second cylinder. 

(c) Referred to third cylinder. 

37. Pressures and percentages used in computing 

the steam accounted for by the indicator 
diagrams, measured to points on the expan- 
sion and compression curves . 

Pressures above zero in pounds per square inch: 

(a) Near cut-off 

(6) Near release 

(c) Near beginning of compression 

Percentage of stroke at points where pressures 
are measured: 

(a) Near cut-off 

(6) Near release 

(c) Near beginning of compression 

38. Aggregate M.E.P. in pounds per square inch 

referred to each cylinder given in heading 

39. Mean back pressure above zero, pounds per 

square inch 

40. Steam accounted for in pounds per indicated 

horse power per hour: 

(a) Near cut-off 

(6) Near release 

41. Ratio of expansion: 

(a) Commercial 

(6) Ideal 



Speed. 

42. Revolutions per minute revolutions. 

43. Piston speed per minute feet. 



Power. 
44. Indicated horse power developed by main-engine cylinders: 

First cylinder horse power. 

Second cylinder 

Third cylinder 

Total " 









APPENDIX C 873 

Efficiency Results. 

45. Dry steam consumed per indicated horse power per hour: 

(a) Main cylinder, including jackets pounds. 

(6) Auxiliary cylinders, etc " 

(c) Engine and auxiliaries 

46. Percentage of steam used by main-engine cylinders accounted for 

by indicator diagrams : 

IstCyl. 2dCyl. 3d Cyl. 
(a) Near cut-off 

(6) Near release 

Sample Diagrams. 

Copies of indicator diagrams, nearest the mean, with corresponding scales, should 
be given in connection with table. 



874 



STEAM POWER PLANT ENGINEERING 



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876 



STEAM POWER PLANT ENGINEERING 






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N 00 00 
LO CO N 
"* -* -* 


O) OS O 
00 OS CN 
■*">*' >o 


t-H CO LO 
** OS ■* 
LO to co 


o o o 


o o o 


o o o 


o o o 


O O O 


O O O 


o o o 


11 
II 


=0 


CN O TtH 

O OS 00 


CO LO CN 

lONO 
N CO CO 


CO OO CO 
CO CD O 
LO * * 


CO O N 

-* os co 

CO CN CN 


N O0 t— I 
00 CO OS 

i-l r-H O 


CO ■* * 
■* CD CN 

O O OS 


O CO T-H 

LO 00 LO 
OO CO LO 


MM(N 


CN CN CN 


CN CN CN 


CN CN CN 


CN CN CN 


CN CN i-H 


T-, ,-H T-H 


o ^ 
H 


+ 


CN -* OS 
OS CD CO 
CO CD CO 


LO O N 

H05CO 
CO LO LO 
LO LO lO 


CO O OO 
-tfl CN OS 
LO LO TtH 
LO LO LO 


CO CD CO 
N LO CO 

lo to to 


CO 00 OS 
t-h OS N 

-* co co 

LO LO LO 


i-H t-H OS 

CO ■* o 

CO CO CO 

»0 LO LO 


CO OS OS 
N OS CN 

CN T-H T-H 
LO LO LO 






Entropy 
of the 
Vapor. 


s.|h 


O OS T-H 

lOOOM 

LO •* * 

o o o 


CO i-H 00 
N CN CO 

CO CO CN 

o o o 


0215 
0164 
0114 


0066 
0019 
9973 


O0 LO .— '. 
MOC* 
OS 00 OO 

Os os Os 


OS OO CO 
OJiON 

n n co 

OS OS OS 


9600 
9419 
9251 










o o o 


o o o 


o o o 










Entropy 

of the 

Liquid. 


<fc 


MiCOO 

Tt< N O 

r-t t-H d 


OS OS OS 

CO CO OS 
cn cn cn 

LO LO LO 


OO CO -* 
CM LO 00 

CO CO CO 

LO LO LO 


ONM 

T-H CO CO 
■* ^ Tt< 
LO LO *C 


00 CO OO 

oo i— i co 

■* LO LO 

to to LO 


CN CD CO 
CO 00 CO 
to LO CO 
»0 LO LO 


CO O OO 
NOON 

CON0O 


o o o 


o o o 


o o o 


o o o 


o o o 


o o o 


o o o 


Heat 

Equivalent 

of 

External 

Work. 


S 


CO LO CO 


N00O5 


O t-h CN 


CO ■* LO 


LO CO N 


NOOO 


T-H CO CO 


CO CO CO 
00 00 00 


CO CO CO 
00 OO OO 


00 OO OO 


00 OO 00 


oo OO OO 


•* Tt< LO 

OO OO OO 


IO LO LO 

O0 00 OO 


Heat 

Equivalent 

of 

Internal 

Work. 


Q. 


^HOO 


CO LO Tji 


TH^HT* 


T*H LO CO 


00 O CO 


LO OO -* 


O CN 00 


000 10 
OONN 

t^ t^ t^. 


MHOS 
N N CD 

t^ t^ t^ 


N LO CO 

co co co 

N N t>- 


H05N 
CO LO LO 

N- t^ !>. 


to rji CN 
LO LO LO 
N N N 


O 00 LO 

LO TjH ■* 

N N N 


CN -* CO 
■* CO CN 
N- t^ N- 




T 

II 
<< 


rfOW 


O -* OS 


"* OO CO 


NHlfl 


00 CN CO 


OS CN OS 


LO O0 T-H 


CO •* rji 

05 0)05 


LO LO LO 

OS OS OS 


CO CO N 

os os os 


N O0 OO 
OS OS OS 


00 OS OS 

os os os 


OS O O 
OS o o 

i-H CN CN 


t-h CN ■* 
O O O 
CN CN CN 




Heat of 
Vapori- 
zation. 


i. 


CN O 00 


CO N N 


OO O0 OS 


ON'* 


CO OS CN 


•* 00 LO 


CO CO CO 


CO i— I 00 
CO CD LO 
00 00 OO 


CO -* CN 

LO LO LO 

00 OO 00 


O OO CO 

LO * * 
O0 00 OO 


LO CO r-H 

OO 00 OO 


OS N CO 

CO CO CO 
OO OO 00 


■* CN OS 
CO CO CN 
00 00 00 


CO 00 i-H 
CN T-H T-H 

OO OO OO 


Heat of 

the 
Liquid. 


O 


CN OS CO 


CN N CN 


(DC* 


NOSrH 


CN •* Tt< 


tO LO T* 


CN CN N 


O CN to 

CO CO CO 
CO CO CO 


OO O CO 
CO CO CO 


LO OO o 

-* TJH LO 

CO CO CO 


CN * N 

LO LO LO 

CO CO CO 


OS t-H CO 
LO CO CO 
CO CO CO 


lO N t-H 

CO CO N 
CO CO CO 


lO * CN 
N 00 OS 
CO CO CO 


Tempera- 
ture, 
Degrees F. 


~ 


LO O CO 


O lO OO 


i-H •* CO 


O0 OS O 


O O OS 


OS 00 "* 


T-H LO LO 


OO i— I CO 
LO CO CO 
CO CO CO 


CD OO O 
CO CO N 
CO CO CO 


CO tO N 

N N N 

CO CO CO 


OS t-H Tfl 
N O0 OO 
CO CO CO 


CO OO OS 
O0 O0 OO 
CO CO CO 


H«N 
OS OS OS 
CO CO CO 


HC5N 

O O t-H 


Absolute 

Pressure, 

Pounds per 

Square 

Inch. 


a, 


O to o 
LO lO CO 


LO O to 
(ONN 


o to o 

OO OO OS 


LO CD LO 
OS O CD 
t-h CN CN 


O to o 

T-H T-H CN 

CN CN CN 


LO o o 

CN CO ■* 
CN CN CN 


CD to O 
tO N O 
CN CN CO 






APPENDIX E. 



EQUIVALENT VALUES OF ELECTRICAL AND MECHANICAL UNITS. 



1 Kilowatt Hour = 
1,000 watt hours 
1.34 horse-power hours 
2,654,200 foot-pounds 
3,600,000 joules 
367,000 kilogram meters 

3.53 pounds of water evaporated 
from and at 212° F. 

1 Kilowatt = 
1,000 watts 

1.34 horse power 
1.358 cheval-vapeur 
2,654,200 foot-pounds per hour 
44,240 foot-pounds per minute 
737.3 foot-pounds per second 
3,412 B.T.U. per hour 

56.9 B.T.U. per minute 
0.948 B.T.U. per second 
3.53 pounds of water evaporated 
from and at 212° F. 

1 Joule = 

1 watt second 

0.000000278 kilowatt hour 
0.102 kilogram meter 

0.0009477 B.T.U. 
0.7373 foot-pound 

1 Watt = 

1 joule per second 

0.00134 horse power 
3.412 B.T.U. per hour 
0.7373 foot-pound per second 
0.0035 pound water evaporated from 
and at 212° F. 

1 Kilogram-Meter = 
7.233 foot-pounds 

0.00000365 horse-power hour 
0.00000272 kilowatt hour 
0.0093 B.T.U. 



1 Horse-Power Hour = 
0.746 kilowatt hours 
1,980,000 foot-pounds 
2,545 B.T.U. 
273,740 kilogram meters 

2.64 pounds of water evaporated 
from and at 212° F. 

1 Horse Power = 
746 watts 
0.746 kilowatts 
1.0136 cheval-vapeur 
33,000 foot-pounds per minute 
550 foot-pounds per second 
2,545 B.T.U. per hour 

42.4 B.T.U. per minute 
0.707 B.T.U. per second 
2.64 pounds of water evaporated 
from and at 212° F. 

1 Foot-Pound = 
1.356 joules 

0.1383 kilogram meter 

0.000000377 kilowatt hour 
0.001285 B.T.U. 
0.0000005 horse-power hour 

1 B.T.U. = 

1,055 watt seconds 
778 foot-pounds 

107.6 kilogram meters 

0.000293 kilowatt hour 

0.000393 horse-power hour 

0.001036 pound water evaporated from 
and at 212° F. 

1 Cheval-Vapeur = 

75 kilogrammeters per second 



0.9863 horse power 
0.7357 kilowatt 



b77 



APPENDIX F. 



MISCELLANEOUS CONVERSION FACTORS. 



1 Pound per Square Inch = 

2.0355 inches of mercury at 32° F. 
2.0416 inches of mercury at 62° F. 
2.309 feet of water at 62° F. 
0.07031 kilogram per square centi- 
meter 
0.06804 atmosphere 
51.7 millimeters of mercury at 

32° F. 

1 Foot of Water at 62° F. = 
0.433 pound per square inch 
62.355 pounds per square foot 
0.883 inch of mercury at 62° F. 
821.2 feet of air at 62° F. and 
barometer 29.92 

1 Inch of Water 62° F. = 

0.0361 pound per square inch 
5.196 pounds per square foot 
0.5776 ounce per square inch 
0.0736 inch of mercury at 62° F. 
68.44 feet of air at 62° F. and 
barometer 29.92 

1 Foot of Air at 32° F. and Barometer 

29.92 = 
0.0761 pound per square foot 
0.0146 inch of water at 62° F. 

1 Inch of Mercury at 62° F. = 
0.4912 pound per square inch 
1.132 feet of water at 62° F. 
13.58 inches of water at 62° F. 



1 Atmosphere = 

760.0 millimeters of mercury at 
32° F. 
14.7 pounds per square inch 
29.921 inches of mercury at 32° F. 
2,116.0 pounds per square foot 

1.033 kilograms per square centi- 
meter 

1 Millimeter = 0.03937 inch 

1 Centimeter = 0.3937 inch 

1 Meter = 39.37 inches 

1 Meter = 32808 feet 

1 Square Meter = 10.764 square feet 

1 Liter = 

61.023 cubic inches 
0.264 U. S. gallons 

1 Gram = 

1 cubic centimeter of distilled 

water 
15.43 grains troy 
0.0353 ounce 

1 Kilogram = 

2.20462 pounds avoirdupois 






APPENDIX G. 

RULES FOR FIREMEN USING ILLINOIS AND INDIANA COAL IN HAND- 
FIRED FURNACES. 

(Formulated by the Coal Stoking and Anti-Smoke Committee of the Illinois 
Coal Operator's Association.) 

1. Break all lumps and do not throw any in furnace any larger than 
one's fist. The reason for this is, that large lumps do not ignite promptly 
and their presence also causes holes to form in the fire, which allow the 
passage of too much air. 

2. Keep the ash pits bright at all times. If they become dark it is 
evident that the fire is getting dirty and needs cleaning, which, if not 
done, will cause imperfect combustion and smoke. If the furnace is 
equipped with a shaking grate, it should be operated often enough to 
prevent any accumulation of ashes in the fire. Do not allow ashes to 
collect in the ash pits, as they not only shut off the air supply, but may 
cause the grate to be burned. 

3. In firing do not land the coal all in one heap, but spread it over 
as wide a space as possible as it leaves the shovel. A little practice 
will enable one to catch the proper motion to give the shovel to make 
the coal spread properly. 

4. Place the fresh coal from the bridge wall forward to the dead plate 
and do not add more than 3 or 4 shovels at a charge. If this amount 
makes smoke it should be reduced till smoke ceases, which means, of 
course, that firing will be at more frequent intervals than formerly to 
keep up steam. This rule applies in cases where the boiler is worked 
at a large capacity. In such instances, however, where a small capacity 
only is required, firing by the coking method is the best, wherein the 
fresh coal is placed at the front of the fire and pushed back and leveled 
when it has become coked. 

5. Fire one side of the furnace at a time so that the other side contain- 
ing a bright fire will ignite the volatile gases from the fresh charge. 

6. Do not allow the fire to burn down dull before charging. If this 
is done, it will not only result in a smoky chimney, but an irregular steam 
pressure. 

7. Do not allow holes to form in the fire. Should one form, fill it by 
leveling and not by a scoop full of coal. Keep the fire even and level 
at all times. As far as possible level the fire after the coal has become 

coked. 

879 



880 STEAM POWER PLANT ENGINEERING 

8. Carry as thick a fire as the draft will allow, but in deciding on the 
proper thickness, judgment must be exercised. If the draft is poor 
a thin fire will be in order, but if strong, a thicker fire should be carried. 

9. Regulate the draft by the bottom or ash pit doors and not by the 
stack dampers, because when the stack damper is used it tends to pro- 
duce a smoky chimney, as it reduces the draft, while the closing of the 
ash pit door diminishes the capacity to burn coal. If strict attention 
is given to firing, and accounting to demand, for steam, there will be no 
occasion to have recourse to dampers, except when there is a sudden 
interruption in the amount of steam being used. 

10. A good general rule is to fire little and often, according to steam 
demands, rather than heavy and seldom. The former means economy 
in fuel and a clean chimney, while the latter signifies extravagance in 
fuel and a smoky chimney. 



APPENDIX H. 



MOLLIER'S DIAGRAM. 



The steam tables give values of the simultaneous physical properties 
of steam, such as pressure, entropy, temperature, etc. When certain 
of these properties are known the remainder can be obtained from the 
tables. The simultaneous properties can also be shown by means of a 
diagram each point on which represents steam in a perfectly definite 
condition. 

Fig. 462 gives a skeleton outline of such a diagram and Fig. 463 a 




H 
G. 462. 



B.T.X7.per Pound of Steam ' ' 2 



reduced reproduction of the complete chart as ordinarily constructed. 
Referring to Fig. 462, abscissas represent the heat contents or B.T.U. 
per pound of steam and ordinates represent the total entropy. Vertical 
lines then represent lines of constant heat content, and horizontal lines 
constant entropy. P 1 P 1 and P 2 P 2 represent lines of constant pressure 
and X X X 1 and X 2 X 2 lines of constant quality. Evidently any point in 
the chart represents a fixed condition of heat content, pressure, quality 
and entropy as determined by its location with respect to the different 
lines. Thus point 1 represents a pressure P t as determined by the 
numerical value of line P 1 P l , quality x x by its location on line X X X U 

881 



882 STEAM POWER PLANT ENGINEERING 

entropy n t by its projection N t on the Y axis, and heat content H l by 
its projection on the X axis. 

The principal advantages of a total heat-entropy diagram over the 
tables are that they give the properties of wet and superheated steam 
and offer a simple means of solving many problems without calculations. 
For example, the chart offers a ready solution of problems involving 

(a) Adiabatic expansion. 

(b) Throttling. 

(c) Expansion with frictional resistances. 

(a) Adiabatic Expansion: From thermodynamics we know that during 
an adiabatic change the entropy is constant; thus, in expanding from 
pressure P 1 and condition represented in point 1 to a lower pressure P 2 
it is only necessary to find the intersection 2 of a horizontal line from 
point 1 with line P 2 P 2 . The various properties corresponding to point 
2 can be read directly from the diagram. 

The line 1-2 = H l H 2 represents the difference in heat content 
following adiabatic expansion from pressure P t and condition 1 to 
pressure P, or 
line #i# 2 =H t — H 2 = x x r x + q t - x 2 r 2 - q 2 . 

The quality x 2 is read directly from the intersection of line 1-2 with 
the constant quality line X 2 X 2 . 

The entropy n 2 , of course, remains the same. 

From equation (73), p. we find that the velocity due to adiabatic 
expansion is 



V = 223.9 V H t - H 



Mollier has added along the margin of the diagram (Fig. 463) a scale 
of velocity so that V may be ascertained by laying off the length H 1 H 2 
on the scale. 

Example: Steam at 120 pounds absolute, quality 0.98, expands adia- 
batically to a back pressure of 2 pounds absolute. Find the quality 
and heat content at the lower pressure. 

From Fig. 462 we locate P x at the intersection of pressure curve 120 
and quality curve 0.98. The corresponding values of H i and n x are 
found by interpolation to be 1174.7 and 1.564 repectively. Follow 
horizontal line 1.564 until it intersects pressure line P 2 . The corre- 
sponding values of H 2 and x 2 are found to be 910 and 0.797 respectively. 
The horizontal intercept between the two pressure lines laid off on the 
velocity diagram gives V = 3640 feet per second. 

Supposing the steam to be superheated 200 degrees instead of being 
wet, find the quality and heat content at the end of expansion. 




— I 



:— I 



E— i 



—I 



884 STEAM POWER PLANT ENGINEERING 

Locate P x at the intersection of pressure curve 120 and superheat 
curve 200. The corresponding values of H 1 and n l are found to be 
1295 and 1.703 respectively. Follow horizontal line 1.703 until it inter- 
sects pressure line P 2 . The corresponding values of H 2 and x 2 are 
found to be 990 and 877 respectively. 

(b) Throttling: If steam expands through a small orifice without the 
addition or abstraction of heat and is brought finally to its initial con- 
dition its total heat will be unchanged. This process is called throttling 
and occurs when steam passes through a reducing valve. Vertical lines 
in Figs. 462 and 463 are lines of constant total heat and consequently 
show the changes in the condition of steam which result from throttling. 
Thus in throttling steam from pressure P 1} Fig. 462, to P 2 it is only 
necessary to find the intersection, 4, of a vertical line from point 1 with 
line P 2 P 2 . 

Example: Steam at 200 pounds pressure and quality 0.96 passes 
through a reducing valve and its pressure is lowered to 15 pounds. 
Find its quality at the lower pressure. The intersection of pressure line 
190 with quality line 0.96 gives H x = 1165. Follow vertical line 1165 
until it intersects pressure line 15. The corresponding value for x 2 
is found to be 30, that is, the steam is superheated 30 degrees. 

To what pressure must the steam be reduced in order that it may be 
dry and saturated? Follow vertical line 1165 until it intersects the 
saturation curve. The corresponding pressure is found to be 30 pounds. 

(c) Expansion Involving Frictional Resistances: As steam expands in 
the nozzle of a turbine or passes between the vanes it experiences fric- 
tional resistances which cause it to give up less energy than it would 
under ideal conditions. The work of friction causes the entropy of the 
steam at its lowest temperature to be greater than it would be if adia- 
batic expansion occurred and serves to increase its dryness fraction. 

If y one hundredths of the heat H 1 — H 2 (given up in adiabatic 
expansion) is lost due to friction, the heat available for useful work is 

{l-y)(H x -H 2 ), 

the resulting velocity of the jet is 

V = 223.9 V(l -y) (H x - H 2 ), 

and the increase in quality of the exhaust steam is 

y(H t -H 2 ) m 

These equations may be readily solved by means of the diagram. 
Referring to Fig. 462 line 1 3 represents an expansion from pressure 
P x to pressure P 2 with frictional resistances. 



APPENDIX H 885 



From the diagram 



y = 



N 2 Hfl 2 



12 HM, 

(1 - y) (H x - H 2 ) = line 1 N = H X H Z . 

Increase in quality = — = distance 2 3 between X 2 X 2 and 

x s x 3 . 

Increase in entropy = NS = N t N 3 . 

Example: Steam at 160 pounds absolute initial pressure, quality 0.97, 
expands through a nozzle to a back pressure of 2 pounds absolute. If 
15 per cent of the heat energy is lost in friction, find the quality of the 
steam at the lower pressure and the velocity of the jet. 

From Fig. 462 we locate P t at the intersection of pressure curve 160 
and quality curve 0.97. The corresponding value of H t is 1170. Follow 
line 1170 horizontally until it intersects pressure line P 2 . From the 
diagram we find for adiabatic expansion # 2 = 910 x 2 = 0.797. But the 
friction increases the heat content at the end of expansion an amount 
0.15 xH t - H 2 = 0.15 (1170 - 910) = 39, so that the final heat con- 
tent = 910 + 39 - 949. 

Follow pressure line P 2 until it intersects heat line 949. The quality 
x 2 is found to be 0.836 and the entropy n 2 = 1.632. From the velocity 
scale we find V = 3320 feet per second for H x - H 2 = (1170 - 949). 



INDEX 



Absorption dynamometer, 857. 

Acetylene, properties of, 26. 

Acidity, tests for, in oils, 674. 

Acme bucket trap, 590. 

Acton atmospheric relief valve, 666. 

Ados C0 2 recorder, 743. 

Aero-pulverizer powdered coal burner, 50. 

Air chambers, 528. 

Air-cooled surface condensers, 428. 

Air lift, 572. 

Air, properties of, 26. 

Air pumps, 552-560. 

size of dry, 558. 

size of wet, 553. 
Air required for operating air lift, 573. 
Air required for combustion, 28. 
Air spaces, grate bars, 114. 
Air supply above grate, 151. 
Air thermometers, recording, 
Air vs. steam as an oil atomizer, 60. 
Alarm, high and low water, 120. 
Alberger barometric condenser, 412. 

cooling tower, 459. 

rotative dry air pumps, 557. 
Alden absorption dynamometer, 857. 
Allis-Chalmers steam turbine, 372. 
Alternate method of starting and stopping 

boiler tests, 826. 
American underfeed stoker, 139. 
Analyses of boiler scales, 473. 

of flue gases, 

of fuel oils, 52. 

of typical American coals, 23. 

of waters for boiler feeding, 473. 
Anchors, pipe, 622. 

Anderson automatic non-return valve, 
658. 

feed-water pumping system, 482. 

triple-duty emergency valve, 659. 
Animal fats and oils, 669. 
Anthracite coals, 15. 



Aqueous vapor, pressure of, 399. 
effect of, on degree of vacuum, 405. 

Armour Glue Works, vacuum ash system 
at, 198. 

Armour Institute, brick chimney at, 227. 

Arndt's econometer, 742. 

Ash bins, 182. 

Ash conveyor, vacuum system, 196. 

Ash, influence of, on fuel value of dry 
coal, 40. 

Ash-handling systems, 181-204. 

Ash, treatment of, in boiler tests, 

A.S.M.E. rules for conducting boiler trials, 
822-845. 

A.S.M.E. rules for conducting engine 
tests, 846-872. 

Atmospheric heaters, 486. 

Atmospheric surface lubrication, 675. 

Atmospheric relief valves, 666. 

Augmenter, Parsons vacuum, 444. 

Aurora and Elgin Interurban Ry., coal- 
handling system, 194. 

Austin steam separator, 579. 

Automatic cut-off vs. throttling engines, 
310. 

Automatic injectors, 547. 

Automatic non-return valves, 658. 

Automatic temperature control, 646. 

Auxiliaries, power consumption of con- 
denser, 449. 

Auxiliaries, measurement of steam used 
by, 851. 

Babcock & Wilcox boilers, 81. 

chain grate, 128. 

superheater, 163. 
Back connection, return tubular boileF, 

79. 
Back pressure on engines, 283. 
Back-pressure valves, 665. 
Baffle-plate steam separator, 579. 



887 



888 



INDEX 



Bagasse as fuel, 20. 
Balanced-draft system, 264. 
Baragwanath feed-water heater, 495. 
siphon condenser, 408. 
surface condenser, 416. 
Barnard- Wheeler cooling tower, 457. 
Basement plan, West Albany station, 

N. Y. C. R. R., 720. 
Bearings, lubrication of, 675. 
Belliss engines, tests of, with superheated 

steam, 314. 
Belt conveyors, 192. 
Bends, pipe, 619. 
Bibliography : 

Cost of electric power, 724. 

Cost of gas power, 726. 

Cost of steam power, 727. 

Cost of water power, 728. 

Description of gas-driven power plants, 

798. 
Description of central stations, steam 

engines, 802-818. 
Description of central stations, steam 

turbines, 808, 819. 
Description of hydraulic power plants, 

798, 818. 
Description of isolated station, 809. 
apartment buildings, 809. 
manufacturing plants, 810. 
office buildings, 812. 
stores, 814. 
Design of power plants, 
Binary- vapor engines, 321. 
Bituminous coals, 16. 
Blades, arrangement of, in steam tur- 
bines, 352, 367. 
Blake jet condenser, 403. 
Blast furnace gas, properties of, 67. 
Bloomsburg steam jet, 246. 
Blowers, fan, 249. 
tests of, 257. 
steam jet, 245. 
Blow-off piping, South Side Elevated 

R.R., 662. 
Blow-offs, 116. 
Blow-off tank, 117. 
Blow-off valve, 661. 
Boiler compounds, 476. 
Boiler-feed pumps {see Pumps). 
Boiler room area, 86. 
Boiler tests, A.S.M.E. code, 822-845. 



Boiler tests, discrepancy between com- 
mercial and experimental results, 842. 
Boilers, 66-123. 

Babcock & Wilcox, chain-grate, 128. 

Babcock & Wilcox, hand-fired,- 80. 

capacity of, 104. 

classification of, 68. 

cost of, 112. 

efficiency of, 98, 762. 

fire-box, 71. 

furnaces for, 68-88, 124-151. 

grates for, 114, 126. 

heating surface of, 92. 

Heine, 83. 

horizontal return tubular, 74. 

horse power of, 93. 

inspector's report (1907), Hartford 
Boiler Insurance Company, 474. 

Manning vertical, 70. 

Parker boilers, 85. 

performances of, 99. 

Robb-Mumford, 73. 

Scotch-marine, 72. 

selection of type, 112. 

settings for, 68-90, 126-140. 

specifications for, 754. 

Stirling, 87. 

vertical tubular, 69. 

Wickes, 84. 
Booth fuel oil burner, 56. 
Boston Elevated, cost of operation, 710. 
Brake horse power, 855. 
Brake, rope, 856. 
Branch fuel oil burner, 57. 
Brass pipes, 608. 
Breeching, 240. 
Brick chimneys, 224. 
Bristol recording air thermometers, 737. 

thermo-electric pyrometer, 737. 
Bucket conveyor, 184. 
Bucket traps, 590. 
Buckeye skimmer, 118. 
Bundy steam separator, 579. 
Bunkers, coal, 182. 
Burgeon, specific heat of superheated 

steam, 157. 
Burke's smokeless furnace, 151. 
Burners, oil, 53-61. 

powdered coal, 44-51. 
Burnham steam meter, 734. . 
Burning point, oils, 674. 



INDEX 



889 



Bursting strength of pipes, 607. 
By-pass system of piping, 625. 

Calorific value of coals, 31. 
Calorimeters, fuel, 747. 
Calorimeters, steam, 745. 
Cannel-coal gas, properties of, 67. 
Capacity, effect of, on boiler efficiency, 104. 
Carbon dioxide, properties of, 26. 

percentage of, in flue gases, 30. 
Carbon monoxide, heat losses due to 
formation of, 36. 

properties of, 26. 
Carbu retted water gas, properties of, 67. 
Carnot cycle, 267. 

Carpenter separating calorimeter, 745. 
Cast-iron pipes, 607. 
Central condensing systems, 439. 
Central hydrostatic cylinder lubricator, 

684. 
Centrifugal oilers, 678. 
Centrifugal pumps, 560-567. 

characteristics of, 565. 

performance of, 563. 

tables of sizes, 566, 567, 

types of, 560. 
Centrifugal steam separators, 578. 
Chain grates, 126. 
Chattanooga Electric Company, test of 

spray fountain, 455. 
Check valves, 660. 

Chemical purification of feed water, 476. 
Chicago setting, hand-fired furnace, 143. 
Chimney at Armour Institute, 227. 
Chimney draft, 207. 
Chimney draft, table of, for various 

temperatures, 210. 
Chimney, efficiency of, 241. 

height of, for burning fuel oil, 218. 

test of 100-foot steel, 213. 
Chimney vs. mechanical draft, 261. 
Chimneys, 207-244. 

brick, 224. 

classification of, 218. 

core and lining for, 230. 

cost of, 243. 

Custodis radial brick, 225, 234. 

dimensions of, 216, 244. 

formulas for, 212. 

foundations for, 240. 

guyed steel, 219. 



Chimneys, materials for brick, 230. 

self-sustaining steel, 219. 

stability of brick, 231. 

stability of steel, 223. 

steel, 219. 

strain sheet for reenforced concrete, 
234. 

thickness of walls, brick, 226. 

thickness of shell, steel, 220. 

Weber reenforced concrete, 235. 
Cincinnati Traction Company, coal con- 
veyors, 195. 
Circulating pumps, 572. 
- Classification of boilers, 68. 

chimneys, 218. 

condensers, 400. 

feed-water heaters, 485. 

fuel-oil burners, 53. 

fuels, 14. 

lubricating oils, 669. 

powdered-coal burners, 44. 

pumps, 522. 

steam separators, 576. 

steam traps, 588. 

steam turbines, 327. 

stokers, 126. 

testing instruments, 730-749. 
Clearance volume, influence of r on engine 

economy, 281. 
Coal, 15. 

analysis of, for boiler tests, 829. 

anthracite, 15. 

bituminous, 16. 

calorific value of, 31. 

composition of, 23. 

measurements of, boiler tests, 852. 

powdered, 44-51. 

proximate analysis, 31. 

purchasing, 42. 

sampling, 828. 

specifications for purchasing, 769. 

storage of, 181. 

ultimate analysis, 31. 

washed, 40. 
Coal and ash handling, 181-204. 
Coal bunkers, 182. 
Coal gas, properties of, 67. 
Coal fields of the United States (ref.), 16. 
Coal hoppers, 202. 
Coal valves, 205. 
Cochrane heater, 487. 



890 



INDEX 



Coefficient of expansion, pipe materials, 

620. 
Coke-oven gas, properties of, 67. 
Cold test, oils, 674. 
Columbia expansion trap, 592. 
Combustion, 24. 
Commercial National Bank Building, 

Chicago, ash system, 192. 
Commonwealth Edison Company, Fisk 

Street Station, 774-787. 
Compressed-air oiling system, 681. 
Compressed air, power required, air lift, 

572. 
Compression, effect of, on engine economy, 

283. 
Compound engines, 300. 
Compounds, boiler, 476. 
Condensers, 397-469. 

air for cooling purposes, surface, 429. 

Alberger barometric, 412. 

Baragwanath siphon, 408. 

barometric, 411. 

Blake jet, 403. 

choice of, 450. 

classification of, 400. 

cooling water for, 404. 

cost of, 450. 

counter-current, 411. 

dry tube, 424. 

economical vacuum for, 451. 

ejector, 410. 

extent of cooling surface for, 421. 

function of, 398. 

high-vacuum, 441. 

independent, 434. 

injection orifice, 407. 

jet, 401. 

Korting multi-jet, 446. 

location of, 433. 

multi-flow surface, 419. 

Schutte ejector, 410. 

siphon, 408. 

sizes of siphon, 409. 

specifications for, 

surface, dry air-cooled, 428. 

surface, evaporative, 433. 

surface, water-cooled, 416. 

tests of surface, 434. 

Tomlinson barometric, 415. 

volume of condenser chamber for jet, 
408. 



Condensers, Weighton multi-flow, 419. 

Weiss barometric, 411. 

Westinghouse-Leblanc, 445. 

Wheeler admiralty, 417. 

Worthington barometric, 414. 
Condensers, Worthington jet, 402. 
Concrete chimneys, 234-240. 
Condensation and leakage losses in en- 
gines, 279. 
Condensing, influence on engine econ- 
omy, 307. 
Condensing plant, elementary, 7. 
Condensing plant with full complement 

of heat-saving appliances, 10. 
Conoidal pump, test of, 568. 
Conversion tables, 878. 
Conveyors, 183-197. 
Cooling ponds, 454. 
Cooling towers, 456. 

Cooling towers, fan vs. natural draft, 460. 
Copper pipes, 607. 
Corliss engine, 269. 
Correction factors for steam turbines, 

387. 
Cost of boilers and settings, 112. 

chimneys, 243. 

condensers, 450. 

engines, 326. 

evaporating water, 105. 

handling coal and ashes, 201. 

mechanical draft systems, 263. 

pipe flanges, 614. 

power (see power costs). 

stokers, 151. 

turbines, 392. 
Costs, operating, 693. 
Coverings for steam pipes, 636. 
Crusher and cross conveyor, 190. 
Curtis steam turbine, 350. 
Curve load factor, 691. 
Custodis radial brick chimney, 234. 
Cut-off, commercial, 860. 
Cut-off, point of, 861. 
Cylinder condensation, 279. 
Cylinder cups, 682. 
Cylinder lubrication, 682. 
Cylinder ratios, compound engines, 300. 

Damper regulators, 118. 
Davis back-pressure valve, 665. 
Dean air pump, 552. 



INDEX 



891 



De Laval centrifugal pump, 568. 

steam turbine, 331. 
Density of air and flue gas, 209. 
Depreciation of powdered-coal furnace,44. 
Depreciation, rate of, 
Depreciation percentages, Chicago Trac- 
tion Valuation Commission, 696. 
Desmond injector, test of, 549. 
Detroit Edison Co., coal-handling system, 

196. 
Diagram factor, steam engine, 862. 
Diaphragm valve, 647. 
Differential traps, 594. 
"Direct" steam separator, 580. 
Disk water meter, 732. 
Divergent nozzle, design of, 334. 
Dodge, A. R., specific heat of super- 
heated steam, 156. 
Double-deck turbine installation, 381. 
Double-flow steam turbine, 369. 
Double stoker, 135. 
Down-draft furnace, 139. 
Draft, balanced, 264. 

chimney, 207. 

for powdered-coal burner, 47. 

forced, 251. 

gauges, 735. 

influence of, on boiler efficiency, 106. 

induced, 252. 

mechanical, 245-266. 
Drainage of jackets and receivers, 597. 
Drains, office building, 603. 
Drips, 586. 

high-pressure, 588. 

low-pressure, 586. 

removal of oil from, 587. 

under alternate pressure and vacuum, 
599. 

under vacuum, 598. 
Dry-air pumps, 537. 
Dry-air surface condensers, 428. 
Dry docks, centrifugal pump character- 
istics for, 565. 
Dry tube surface condensers, 424. 
Dulong's formula, 32. 
Dunham steam trap, 593. 
Duplex coal valve, 204. 
Duplex steam pump, 524. 
Duplicate piping system, 624. 
Dutch oven, 141. 
Duty, pump, 536. 



Economizers, 508. 

factors for determining installation of, 
532. 

Green, 510. 

heat transmission in, 512. 

tests of, 515. 
Edwards air pump, 555. 
Efficiencies of boilers and grates, 98. 

boilers with oil fuel, 54. 
Efficiencies of boiler-feed pumps, 534. 

centrifugal pumps, 565-568. 

compound engines, saturated steam, 
306. 

compound engines, superheated steam, 
317, 321. 

fans, 257. 

piston pumps, 533. 

simple engines, saturated steam, 296. 

simple engines, superheated steam, 316. 

steam turbines, 384. 

triple-expansion engines, superheated 
steam, 318. 

triplex pumps, 544, 545. 
Efficiency, air lift, 574. 

Carnot cycle, 267. 

condensing plants, 9, 13. 

furnace, 99. 

grate, 99. 

mechanical, 275. 

non-condensing plants, 4. 

Rankine cycle, 268. 

thermal, 273. 
Ejector condenser, 410, 446. 
Ejector, Shone, 603. 
Electrical power, cost of, 704-729. 
Elementary condensing plants, 2. 
Elementary non-condensing plant, 7. 
Elementary theory, Curtis turbines, 358. 

De Laval turbine, 333. 

Westinghouse-Parsons turbine, 373. 
Elevating tower, cable-car distribution, 

193. 
Elevating tower, hand-car distribution, 

196. 
Ellison's universal steam calorimeter, 746. 
Emergency valves, 658. 
Engines (steam), 267-326. 

A.S.M.E. code for testing, 846-872. 

automatic cut-off, 310. 

back pressure, effect of, on economy, 
283. 



892 



INDEX 



Engines (steam), binary vapor, 321. 

clearance volume, 281. 

compound, 300. 

compression, effect of, on economy, 
283. 

condensing, effect of, on economy, 307. 

cost of, 326. 

cylinder condensation, 279. 

economy of (see Tests). 

efficiencies of (see Efficiencies). 

friction of, 284. 

heat losses in, 278. 

high-speed, 290. 

ideal, 267. 

incomplete expansion, loss due to, 282. 

increasing initial pressure, effect of, 
286. 

jackets, influence of, 170. 

leakage losses in, 279. 

low-speed, 299. 

mechanical efficiency, 275. 

non-condensing, test of, 296. 

receiver-reheaters, economy of, 287. 

simple, 291. 

single-acting, 290. 

specifications, 750. 

sulzer, 319. 

superheated steam, 313. 

tests of (see Tests). 

thermal efficiency of, 273. 

throttling vs. automatic cut-off, 310. 

triple and quadruple expansion, 305. 

wire drawing, effect of, 284. 

with low-pressure turbines, 376. 
Entropy diagram, 863, 881. 
Equation of pipes, 640. 
Exhaust heads, 585. 
Exhaust piping, 642. 
Expansion traps, 592. 
Expansion of pipe materials, 618. 
Expansion, ratio of, 862. 
Extended boiler front setting, 
Evaporation, cooling pond, 454. 

cost of, coal fuel, 105. 

cost of, oil vs. coal, 55. 

from and at 212° F. per square foot per 
hour, 95. 

rate of, boilers, 96. 

rates of, in still air, 454. 

unit of, 88. 
Evaporative surface condenser, 433. 



Factor of evaporation, 88. 

Fan draft, 249. 

Fans, capacity of induced-draft, 262. 

capacity of forced-draft, 261. 

performance of, 258. 

theory of, 252. 
Feed water, analyses of, 473. 
Feed-water heaters (see Heaters). 
Feed-water heating system, choice of, 516. 
Feed-water piping, 647. 
Feed- water purification, 471. 
Feed-water regulators, 542. 
Fery radiation pyrometer, 740. 
Filters, oil, 688. 
Fire-box, boilers, 71. 
Fire-tile "Economy," 131. 
Fire-tile combustion chamber, 128. 
Fire, thickness of, 110. 
Fire-tube boiler, 71. 
First National Bank Building, Chicago, 

power costs, 710. 
Fisher pump governor, 541. 
Fittings, pipe, 610. 
Fixed carbon in coal 18. 
Fixed charges, 693. 
Flanged fittings, 610. 
Flanges, table of extra heavy, 615. 
Flanges, table of standard, 614. 
Flap coal valve, 205. 
Flash point, oil testing, 673. 
Flemming four-valve engine, 320. 
Flinn trap, 595. 
Float trap, 589. 
Floor space, turbine vs. reciprocating 

engine, 383. 
Flow of steam in pipes, 632. 

steam through nozzles, 336. 

water through pipes, 650. 
Flue-gas analysis, 29, 741. 
Flue-gas apparatus, Ados recording, 743. 

Arndt's econometer, indicating, 742. 

Orsat, 741. 

Sarco recording, 744. 
Flush-front boiler setting, 
Fly-wheel pumps, 523. 
Foot valves, 668. 
Forced draft, 245-266. 
Forced-feed lubricator, 684. 
Forcing capacity of boilers, 109. 
Foster back-pressure valve, 665. 

pressure regulator, 667. 









INDEX 



893 



Foster superheater, 165. 

Foundations, chimney, 240. 

Fountain, spray, 455. 

Four- valve engines, tests of, 298, 306. 

Friction of engines, 284. 

Friction of water in pipes, 652. 

Friction tests of oil, 674. 

Friction through valves and fittings, 639, 
653. 

Fuel, cost of, 700. 

Fuel calorimeters, 747. 

Fuel oil, 51-66. 

Fuel-oil burners (see Burners). 

Fuels and combustion, 14-16. 

Fuels, classification of, 14. 

Function of the condenser, 398. 

Furnace arch bars, 79. 

Furnace efficiency, 99. 

Furnace temperature, influence on boiler 
efficiency, 111. 

Furnace influence on gas composition, 36. 

Furnace for burning oil fuel, 59. 

Furnace for burning powdered coal, 47. 

Furnace, smokeless (see Smokeless fur- 
naces). 

Fusible plugs, 121. 

Gaseous fuels, 66. 

characteristics of, 67. 
Gauge cocks, 3, 119. 
Gauges, water, 119. 
Gate valves, 655. 
Geipel steam trap, 593. 
Globe valves, 655. 
Government specifications for purchasing 

coal, 769. 
Governor, steam pump, 541. 
Goubert feed-water heater, 492. 
Grate, loss of fuel through, 37. 
Grate surface, 95, 98. 
Grate bars, thickness of, 114. 
Grates, chain, 126. 

rocking, 116. 

stationary, 114. 
Gravity oil feed, 680. 
Gravity, Baume oils, 672. 
Gravity, specific oils, 672. 
Grease extractor, 584. 
Greases, 671. 
Green chain grate, 126. 

economizer, 510. 
Guyed steel chimneys, 219. 



Hamilton-Holzworth turbine, 362. 

Hamler-Eddy Smoke Recorder, 

Hammel fuel oil burner, 57. 

Hancock injector, 546. 

Hand shoveling, 183. 

Hangers, for pipes, 622. 

Hartford Boiler Insurance Company, 

annual report (1907), 474. 
Hartford boiler specifications, 754. 
Hawley down-draft furnace, 139. 
Headers, main steam, 563. 
Heat balance, boiler tests, 832. 
Heat distribution, condensing plants, 9, 

13. 
Heat distribution, non-condensing plants, 

4. 
Heat losses in burning coal, 32. 

in the chimney gases, 33. 

in steam engines, 278. 
Heat transmission, boilers, 90, 

closed heaters, 497. 

economizers, 512. 

influence of scale on, 472. 

superheaters, 170. 
Heating surface, boilers, 92, 95. 
Heaters, feed- water, 471-521. 

Baragwanath, 595. 

choice of, 516. 

classification of, 485. 

closed, 486. 

Cochrane, 487. 

counter-current, 485. 

flue-gas, 485. 

Goubert, 492. 

Harrisburg, 494. 

Hoppe's, 489. 

induced, 486, 506. 

live steam, 485-507. 

open, 486, 504. 

Otis, 494. 

parallel-current, 492. 

primary, 486. 

secondary, 486. 

single-flow, 492. 

steam tube, 494. 

through, 486, 505. 

vacuum, 485, 506. 

Wainwright, 493. 

Webster, 488. 
Heater and purifier combined, 489. 
Heine boiler, 83. 



894 



INDEX 



Heinrich smokeless furnace, 249. 

Heintz expansion trap, 594. 

Herringbone grate, 115. 

Hewes and Phillips air pump, 556. 

Heyworth Building, Chicago, plan of 
piping, 421. 

High and low speed engines, 290. 

High-pressure drips, 588. 

High-speed double- valve engine, 297. 

High-speed single- valve engine, 290. 

Hollow bridge wal, 151, 247. 

Holly loop, 602. 

Hoppe's feed-water heater, 489. 
steam separators, 577. 

Horizontal return tubular boilers, 74. 

Horse power of boilers, 93. 

Hot-well temperatures, surface con- 
densers, 426. 

Hot-well pumps, 560. 

Hunt coal conveyor, 189. 

Hydraulic packing, 530. 

Hydraulic Oil Storage Company's fuel 
oil system, 64. 

Hydraulic valve gear, Curtis turbine, 355. 

Hydrogen, properties of, 26. 
losses due to, 38. 

Hydrometer, Baume, oils, 672. 

Hydrostatic cylinder lubricator, 683 

-Hygrometry, 468. 

Ideal engine, 267. 

Illinois Engineering Company's auto- 
matic vacuum valve, 643. 

Impulse turbine, 331. 

Incomplete combustion, loss due to, 36. 

Incomplete expansion, loss due to, 282. 

Increasing boiler pressure, economy of, 
286. 

Increasing rotative speed, economy of, 
290. 

Increasing degree of vacuum, cost of, 451 . 

Increasing degree of vacuum, economy 
of, 308, 395. 

Identification of oils, 672. 

Independent condensers, 434. 

Independently fired superheaters, 166. 

Indicated horse power, 853. 

Indicator cards, air pump, 560. 

Indicator cards, analysis of, 859. 
automatic cut-off engines, 311. 
four- valve engine, 861. 



Indicator cards, throttling engine, 311. 

Westinghouse-Parsons turbine, 365. 
Indicator springs, tests of, 854. 
Induced draft, 252. 
Induced heaters, 506. 
Initial condensation, 279. 
Injection orifice, 408. 
Injectors, 545-548. 

performance of, 548. 

range in working pressures, 350. 

vs. steam as boiler feeders, 550. 
Intermittent oiling, 675. 
International Gas Company's fuel oil 

system, 63. 
Interest charges, 693. 
Isolated stations, cost of power in, 71 6-723. 
Isolated stations, influence of load factor 
on economy, 720. 

Jackets, influence of, 170. 

Jackets, methods of draining, 597. 

Jet condensers, 

Jet, Bloomsburg, 246. 

Jet, ring steam, 246. 

Jets, steam consumption of, 248. 

Jones underfeed stoker, 72. 

Kent's wing-wall furnace, 149. 

Kerosene, use of, in boilers, 477. 

Kerr steam turbine, 346. 

Keystone separator, 578. 

Kieley reducing valve, 667. 

Kindling temperatures, 25. 

Kirkwood oil burner, 58. 

Kitts feed-water regulator, 542. 

Kitts hydraulic damper regulator, 53. 

Knoblauch and Linde, specific heat of 

superheated steam, 156. 
Knowles triplex pump, test of, 544. 
Korting fuel-oil burner, 56. 
Korting multi-jet condenser, 446. 

Labor, cost of, in power plants, 699. 

cost of, in street railway plants, 701. 

cost of, in tall office buildings, 702. 
Lea-Degan three-stage turbine pump, 562, 

569. 
Leakage of steam in engines, 279. 
Leyland automatic cylinder cup, 683. 
Life of power plant appliances, 694. 
Lignite, 18. 
Limit of superheat, 154. 



INDEX 



895 



Link Belt Company coal-handling sys- 
tem, 184. 
Live-steam feed-water heaters, 507. 
Load factor, 691. 

influence of, on cost of power, 692. 
Location of condensers, 433. 

of separators, 580. 

of traps, 596. 
Loew grease extractor, 584. 
Loop header, 624. 
Loop, Holly, 602. 
Loop, steam, 600. 
Loss of heat from bare pipes, 615. 
Loss of heat from covered pipes, 616. 
Losses in burning fuel, 32-40. 
Losses in steam engines, 278. 
Low-pressure drips, 586. 
Low-pressure turbines, 376. 
Low-speed engines, 299. 
Lubricants, 669-680. 
Lubrication, atmospheric, 675. 

central hydrostatic system, 684. 

compressed-air feed, 680. 

cylinder lubrication, 682. 

forced system, 684. 

gravity systems, 680. 

Siegrist system, 685. 
Lubricating oils, classification of, 670. 

properties of, 676. 

specific gravity of, 672. 
Lubricators, hydrostatic, 683. 
Ludlow angle valve, 667. 
Lunkenheimer sight-feed lubricator, 684. 

Mahler bomb calorimeter, 747. 

Mains, steam, 622. 

Maintenance, 699. 

Manning vertical boilers, 70. 

Marks' and Davis' steam tables, 

Marsh gas, properties of, 26. 

Marsh steam pump, 527. 

Marsh steam pump, test of, 534. 

Materials for brick chimneys, 230. 

Materials for pipes and fittings, 606. 

Materials for superheaters, 170. 

McClaves argand blower, 246. 

McDaniel float trap, 589. 

Mean temperature difference, heaters, 

497. 
Measurement of heat units consumed by 

engines, 849. 



Measurement of feed-water consumption, 

engine, 849. 
Measurement of steam used by auxiliaries, 

850. 

Mechanical draft, 245-266. 
Mechanical efficiency of engines, 275. 

of pumps, 533. 
Mechanical boiler-tube cleaner, 121. 
Mechanical purification of feed water, 481. 
Mechanical stokers, 125. 
Mechanical valve gear, Curtis turbine, 

354. 
Medium and low speed engines, 299. 
Mesh steam separators, 580., 
Meters, steam, 734. 
Meters, water, 732. 
Meyers Bagasse furnace, 21. 
Meyers tan bark furnace, 22. 
Mineral oils, 670. 
Mixed pressure turbines, 376. 
Moisture in air, loss due to, in combustion, 
37. 

in fuel, loss due to, 37. 

in steam, determination of, 827. 

in steam, effect on engine economy, 
278. 

in steam evaporated by throttling, 312. 
Mollier diagram, 
Mullan valveless air pump, 556. 
Murgue's theory, centrifugal fans, 254. 
Murphy furnace, 137. 

Napier's rule for the flow of steam, 336. 
Natural-draft cooling tower, 460. 
Natural gas, properties of, 67. 
Naval Liquid Fuel Board, report of 

United States, 65. 
Nitrogen, properties of, 26. 
Non-condensing engines, test of, 296, 315. 
Non-condensing plants, arrangement of, 
2,5. 
exhaust piping in, Paul system, 643. 
exhaust piping in, Webster system, 542. 
feed- water piping in, 647. 
open heater in, 506. 
Non-return valves, 658. 
Norfolk Traction Co.'s ash-hananiig sys- 
tem, 200. 
Northwestern Elevated R. R. power 

house, condenser piping, 446. 
Nozzles, De Laval steam turbine, 332. 



896 



INDEX ' 



Nozzles, flow of steam through, 334. 

flow of water through, 650. 

Kerr steam turbine, 346. 

theoretical design of divergent, 334. 
Nugent telescopic oiler, 677. 

Office buildings, cost of power in, 703- 

720. 
Oil bath, 675. 
Oil burners (see Burners). 

cups, 677. 

eliminators, 581. 

filters, 687. 

fuels, analysis of, 31. 

pressure in fuel oil systems, 63. 

separators, 581. 

storage, 64. 
Oil, waste, and supplies, cost of, in power 

plants, 703. 
Oiler, centrifugal, 678. 

gravity, 679. 

pendulum, 678. 

ring, 678. 

telescope, 677. 
Oiling systems (see Lubricating systems). 
Oils, animal, 669. 

distilled, 670. 

identification of, 672. 

mineral, 670. 

properties of, 671. 

specific gravity of, 672. 

specifications for, 672. 

tests for, 672. 
Olefiant gas, properties of, 26. 
Open heaters, 486. 
Open heater vs. closed heater, 504. 
Operating costs, 
Operating costs, reciprocating engine vs. 

steam turbine, 392. 
Optical pyrometers, 738. 
Orsat apparatus, 741. 
Orifice, size of injection, 408. 
Orifices, flow of steam through, 336. 

flow of water through, 584. 
Otis feed-water heater, 428. 
Overhead storage, bucket hoist, 195. 
Overload capacity, steam turbines, 364. 
Oxygen, properties of, 26. 

Pan surface, open heaters, 491. 
Parallel-current condenser, 401. 
Parallel-current feed-water heater, 492. 



Parker boiler, 85. 

Parr coal calorimeter, 748. 

Parsons smokeless furnace, 248. 

Parsons vacuum augmenter, 444. 

Paul exhauster, 645. 

Paul heating system, 643. 

Peat, 18. 

Penberthy injector, 546. 

Pendulum oiler, 678. 

Pennel saturated-air surface condenser, 

431. 
Pinther powdered-coal burner, 49. 
Pipe anchors, 622. 
Pipe bends, 619. 
Pipe, brass, 608. 

cast-iron, 607. 

cast-steel, 607. 

copper, 608. 

mild steel, 607. 

sizes of standard, 611. 
Pipe flanges, 610. 

size of extra heavy, 615. 

sizes of standard, 614. 
Pipe hangers, 622. 
Pipe supports, 621. 

Pipe threads, United States standard, 615. 
Pipes, equation of, 636. 

flow of steam in, 632. 

flow of water in, 650. 

size of, for low-pressure drips, 588. 

strength of, 609. 
Piping, by-pass system of steam, 625. 

Commonwealth Edison Company, Fisk 
Street Station, 774-787. 

condenser, 642. 

Des Moines City Railway Company, 
631. 

duplicate system of, 624. 

feed-water, 648. 

Heyworth Building, Chicago, 626. 

loop header system of, 624. 

Manhattan Elevated, New York, 627. 

Paul heating system, 643. 

Princeton University, 623. ' 

specifications for, 760. 

steam, 622. 

Webster heating system, 641. 

West Albany Station, New York Cen- 
tral Railway Company, 788. 
Pistons, water, 530. 
Pitot tubes, 253. 









INDEX 



897 



Plungers, pump, 530. 
Ponds, cooling, 454. 
Pop safety valves, 664. 
Positive injectors, 547. 
Powdered coal, 44. 
Powdered-coal burners (see Burners). 
Power consumption of condenser auxil- 
iaries, 449. 
Power cost, Boston Elevated, 710. 

British electric light and power plants, 
709. 

compound engine plants, 711. 

depreciation, 694. 

First National Bank Building, Chi- 
cago, 710-724. 

fixed charges, 693. 

fuel, 700. 

insurance, 699. 

interest, 699. 

isolated stations, 723. 

labor, 699. 

maintenance, 699. 

operating charges, 699. 

simple engine plant, 711. 

street railway plants, C. C. Moore, 
701-705. 

street railway plants, R. C. Carpenter, 
715. 

street railway plants, typical U.S., 709. 

taxes, 699. 
Power measurement, 741. 
Powers thermostat, 646. 
Preheating feed water, economy of, 484. 
Pressure gauges, 735. 
Pressure of aqueous vapor for different 

temperatures, 399. 
Pressure regulation, 667. 
Producer gas, properties of, 67. 
Properties of air, 462. 

of fuel oil, 52. 

of gases, 57. 

of lubricating oils, 670, 676. 

of steam, 873. 
Proximate analysis, 31. 
Pulsometer, 572. 
Pump governors, 541. 
Pumping engines, surface condensing for, 

436. 
Pumps, air, 552-560. 

air lift, 572. 

air, jet condensers, 552. 



Pumps, air, sizes of, 553. 

air, surface condensers, 558. 

air, theoretical work, 560. 

boiler-feed, 524. 

classification of, 522. 

centrifugal, 522, 523, 560-566. 

circulating, 572. 

direct-pressure, 528. 

duplex, 524. 

duty of, 536. 

effect of piston speed on economy, 473. 

fly-wheel, 532, 561. 

jet, 522, 523, 566. 

Marsh boiler-feed, 527. 

multi-stage centrifugal, 561. 

outside packed plunger, 531. 

performance of piston, 531. 

power, 543. 

rotary, 567. 

simplex, 537. 

size of boiler-feed, 539. 

tests of (see Tests). 

triplex, 543. 

turbine, 561. 

volute, 561. 

water pistons for, 530. 
Purchasing coal, 42. 

government specifications for, 769. 
Purification, chemical, feed water, 476. 

mechanical, 479. 

thermal, 479. 
Purifiers, live-steam, 507. 
Purifying plants, Anderson, 583. 

Scaife, 581. 

We-Fu-Go, 581. 
Pyrometers, air-recording, 737. 

Bristol thermo-electric, 737. 

Callendar resistance, 738. 

Fery radiation, 740. 

Wanner optical, 738. 

Quality of steam, 827. 

Qualifications for a good lubricant, 671. 

Radial brick chimneys, 234. 
Radiation and minor losses, boilers, 39. 
Radiation pyrometer, 740. 
Rankine cycle, 268. 

Rate of combustion, powdered coal, 45. 
Rate of combustion, relation of air sup- 
ply on, 264. 



898 



INDEX 



Rate of depreciation, 694. 

Rate of driving, effect on economy of 

boilers, 106. 
Rateau regenerator accumulator, 380. 
six-stage turbine pump, 562. 
steam turbine, low-pressure, 376. 
Ratio of cooling water to condensed 

steam, 406. 
Ratio of heating to grate surface, 98. 
Ratio of expansion, 862. 
Reaction turbines, 365. 
Receiver-reheaters, 287. 
Reciprocating engines vs. steam turbines, 

386. 
Records, power plant, 690. 
Reducing valves, 396, 462. 
Reenforced concrete chimneys, 234. 
Regulation of steam turbines, 353, 368. 
Regulators, damper, 118. 

feed- water, 541. 
Repairs, cost of, power plants, 703. 
Report of United States Naval Liquid 

Fuel Board, 65. 
Restricted feed, lubrication, 675. 
Returns tank, 603. 
Ring oiler, 678. 
Ring steam jet, 246. 
Ringleman smoke chart, 844. 
Riveted joints, steel chimneys, 223. 
Robb-Mumford boiler, 73. 
Robins belt conveyor, 192. 
Rochester forced-feed lubricator, 685. 
Roney stoker, 134. 
Rope brake, 856. 
Rotary pumps, 523, 567. 
Rowe feed-water regulator, 542. 

Safety valve, dead weight, 663. 

level, 563. 

pop, 564. 
Safety valves, capacity of, 565. 

rules for loading, 564. 
Sampling coal, 828. 
Sarco C0 2 recorder, 744. 
Saturated-air surface condenser, 430. 
Saturated-steam tables, 893. 
Sawdust as fuel, 18. 
Scaife system of feed-water purification, 

481. 
Scale, analyses of boiler, 473. 

influence on heat transmission, 472. 



Schmidt, independently fired superheaters, 

167. 
Schutte ejector condenser, 410. 
Schwartzkopff powdered-fuel burner, 50. 
Scotch marine boiler, 72. 
Screwed fittings, 610. 
Seaton coal valve, 205. 
Separating calorimeters, 745. 
Separators, 575-586. 

Austin, 579. 

baffle-plate, 579. 

Baum, 583. 

Bundy, 579. 

centrifugal, 578. 

classification of, 576. 

direct, 580. 

exhaust steam, 581. 

Hoppes, 577. 

Keystone, 578. 

live steam, 576. 

location of, 580. 

mesh, 579. 

oil, tests of, 582. 

reverse current, 577. 

Stratton, 578. 
Settings, smokeless, 124- 

standard, 68- 
Shone ejector, 604. 
Siegrist oiling system, 686. 
Simple engines, 291. 
Simplex coal valve, 204. 
Single vs. double-acting engines, 291. 
Siphon condensers, 408. 
Siphon traps, 595. 
Skimmer, Buckeye, 118. 
Slip expansion point, 621. 
Smoke observation, 844. 
Smoke prevention, 124- 
Smoke chart, 844. 
Smoke recorders, 749. 
Smoke, visible, loss due to, 32. 
Smokeless furnaces, 124- 

balanced draft, 264. 

Burke's, 151. 

Chicago Smoke Inspection Depart- 
ments, 143. 

Dutch oven, 141. 

fire tile, 128. 

Heinrich's, 249. 

Kent's wing wall, 149. 

Parsons, 248. 



INDEX 



899 



Smokeless steam jets, 245-250. 

stokers, 125-140. 
*" twin fire-arch, 143. 

Wooley, 148. 
Solid lubricants, 671. 

South Side Elevated Ry. Co., Chicago, 
chimney at, 221. 

coal crusher and cross conveyor at, 190. 
Special furnaces, 141. 
Specific heat of superheated steam, 155- 
162. 

A. R. Dodge, 156. 

C. C. Thomas, 161. 

C. E. Burgoon, 157. 

Knoblauch and Jakob, 156, 159. 

value of c p at atmospheric pressure, 
157. 
Specific volume of superheated steam, 160. 
Specific gravity of lubricating oils, 672. 
Specifications, boiler, 754. 

condenser, 758. 

engine, 750. 

government, for purchasing coal, 769. 

piping, 760. 
Speed, influence on piston pump econ- 
omy, 2. 

influence on engine economy, 290. 
Split bridge wall, 151. 
Spray fountain, 455. 
Sprinkling stokers, 141. 
Stability of brick chimneys, 231. 
Stability of steel chimneys, 223. 
Standard method of starting and stop- 
ping boiler tests, 825. 
Station load factor, 691. 
Steam boilers, 68. 
Steam engines, 267- 
Steam, flow of, in pipes, 632. 
Steam jets, 245-250. 
Steam loop, 600. 
Steam mains, 629. 
Steam piping, 606, 668. 
Steam, properties of, 878. 
Steam pumps, 522-541. 
Steam separators, 575-605. 
Steam, specific heat of superheated, 130- 

155. 
Steam traps, 575-605 
Steam turbines, 327. 
Steel concrete chimneys, 234. 
Steel chimneys, 219. 



Step bearing, 356. 
Stirling boiler, 87. 
Stirling superheater, 164. 
St. John's steam meter, 734. 
Stokers, 125-140. 

American underfeed, 139. 

Babcock & Wilcox, 128, 132. 

chain grates, 126. 

cost of, 151. 

Green chain grate, 127. 

Jones underfeed, 138. 

Murphy, 137. 

Roney, 134. 

Wilkinson, 136. 
Stop valves, 655. 
Storage, coal, 181. 

oil, 64. 

powdered coal, 45. 
Stratton separator, 578. 
Sulphur in fuel, 25, 40. 
Sulphur dioxide, properties of, 26. 
Sulzer engine for superheated steam, 

319. 
Superheat, limit of, 154. 
Superheated steam, 153-180. 

economy of, steam engine, 313. 

economy of, steam turbines, 386. 

properties of, 180. 

specific heat of, 155. 

specific volume of, 153, 160. 
Superheating moisture in air, loss due 

to, 37. 
Superheating surface, extent of, 170. 
Superheaters, 163- 

Babcock & Wilcox, 163. 

Foster, 165, 168. 

independently fired, 166. 

Schmidt, 167. 

Schwoerer, 166. 

Stirling, 164. 

tests of independently fired vs. flue 
fired, 176. 

tests, miscellaneous, 176-179. 

yearly expense for repairs, cast-iron, 
171. 
Supports, pipe, 621. 
Surface blow, 117. 
Surface condensers, 416-430. 
Surface, cooling, condensers, 421. 
Surface, heating, feed-water heaters, 496. 
Surface, heating, superheaters, 170. 



900 



INDEX 



Tan bark as fuel, 18. 
Tank, blow-off, 117. 
Tank, returns, 602. 
Taxes, power cost, 693. 
Telescopic oiler, 677. 
Temperature-entropy diagram, 864, 881. 
Temperature of combustion, 27. 
Temperature measurements, 736. 
Temperature regulators, 646. 
Terry steam turbine, 345. 
Tests of blowers, 257. 
Tests of boilers: 

A.S.M.E. code, 822. 

coal burning, 100. 

evaporation, Armour Glue Works, 105. 

influence of draft on efficiency, 106. 

influence of rate of combustion on air 
supply, 264. 

influence of rate of driving on capac- 
ity, 107. 

influence of size of coal on capacity, 40. 

influence of thickness of fire on effi- 
ciency, 109. 

oil fuel, 62. 

powdered coal, 46. 
Tests of burners, oil fuel, 62. 
Tests of chimney, 100-foot steel, 213. 
Tests of condenser auxiliaries, 449. 
Tests of condensers: 

dry air, 429. 

evaporative, 434. 

Pennel saturated air, 432. 

Weighton surface, 421. 
Tests of cooling towers, 468. 
Tests of economizers, 515. 
Tests of engines, A.S.M.E. code, 846. 

binary vapor, 324. 

compound condensing vs. non-con- 
densing, 304. 

compound vs. simple high-speed, 303. 

condensing, increase in power due to, 
307. 

Corliss, 369. 

5500-h.p. engine at Waterside Station, 
309. 

four-valve vs. single-valve high-speed, 
298. 

friction, 284. 

increasing back pressure, 283. 

increasing initial pressure, 286. 

Reeves simple engine, 297. 



Tests of engines, simple high-speed engines, 
saturated steam, 292. 

tables of, 296, 306, 

superheated steam, compound, 317. 

superheated steam, influence of degree 
of superheat, 322. 

superheated steam, influence of size of 
engine, 332. 

superheated steam, record perform- 
ance of, 321. 

superheated steam, triple, 318. 
Tests of furnace, relation of gas compo- 
sition to temperature, 36. 
Tests of injectors, 549. 
Tests of jets, steam, 248. 
Tests of oil burners, 62. 
Tests of oil separators, 572. 
Tests of oils, fuel, 52. 
Tests of oils, lubricating, 676. 
Tests of pipe coverings, 617. 
Tests of pumps: 

air lift, 572. 

boiler feed, 534. 

centrifugal, De Laval, 568. 

centrifugal, Lea-Degan, 569. 

centrifugal, Morris, 564. 

centrifugal, Worthington, 568. 

direct-connected triplex, 544. 

duplex fire pump, 533. 

geared triplex, 545. 

rotary, 570. 
Tests of separators, 576. 
Tests of spray fountain, 455. 
Tests of superheaters, 176-181. 
Tests of turbines, 390. 
Thermal efficiency of engines, 273. 
Thermal purification of feed water, 479. 
Thermo-electric pyrometers, 736. 
Thermometers, classification of, 731. 
Thermostat, Powers, 646. 
Thickness of fire, 109. 
Thickness of walls, brick chimneys, 226. 
Thomas, specific heat of superheated 

steam, 161. 
Throttling, calorimeter, 746. 
Throttling, moisture evaporated by, 312. 
Throttling vs. automatic cut-off, 310. 
Tile, "Economy," 131. 
Tile-roof furnaces, 63. 
Tilden damper regulator, 119. 
Tomlinson condenser, 415. 



INDEX 



901 



Towers, cooling, 456-470. 

Alberger, 459. 

Barnard, 457. 

test of, 468. 

theory of, 460. 

Worthington, 458. 
Traps (steam), 588-599. 

Acme, 590. 

bowl, 591. 

bucket, 590. 

Bundy, 591. 

classification of, 588. 

Columbia, 592. 

differential, 594. 

Dunham, 593. 

expansion, 592. 

Flinn, 595. 

float, 589. 

Geipel, 593. 

Heintz, 594. 

location of, 596. 

McDaniel, 589. 

return, 596. 
Traveling coal hoppers, 203. 
Traveling grates, 126. 
Triumph powdered-coal furnace, 50. 
Try cocks, 3, 120. 
Tube cleaners, 121. 
Tupper grate bar, 115. 
Turbines (steam), 327. 

advantages of, 382. 

Allis-Chalmers, 372. 

correction factors, 387. 

cost of, 392. 

Curtis, 350. 

De Laval, 331. 

Double-flow, 369. 

economy of space, 382. 

efficiency of, 384. 

elementary theory, 328. 

Hamilton-Holzworth, 362. 

impulse, 331. 

influence of superheat, 393. 

influence of vacuum, 395. 

Kerr, 346. 

low-pressure, 376. 

overload capacity, 384. 

reaction, 365. 

regulation, 384. 

simplicity of, 382. 

Terry, 345. 



Turbines (Steam), tests of, 390. 

Westinghouse-Parsons, 365. 
Turner oil filter, 688. 
Twin fire-arch furnace, 142. 

Ultimate analysis, 31. 

Underfeed stokers, 138. 

Unit of evaporation, 88. 

Units, conversion, 877. 

Universal calorimeter, 746. 

Useful life of power-plant appliances, 694. 

Vacua, increase of power due to increas- 
ing, 307. 
Vacua, influence of, on economy of engines, 

307. 
Vacua, influence of high, on steam tur- 
bines, 395. 
Vacuum ash conveyor, 196. 
Vacuum augmenter, Parsons, 444. 
Vacuum chambers, 529. 
Vacuum, degree of, as affected by aqueous 
vapor, 405. 

drips under, 597. 

most economical, 451. 
Vacuum pumps, 552. 
Vacuum systems, high, 441. 
Valves, Act on atmospheric relief, 666. 

Anderson automatic non-return, 658. 

Anderson triple duty emergency, 659. 

atmospheric relief, 665. 

back pressure, 665. 

blow-off, 661. 

by-pass, Westinghouse-Parsons, 369. 

check, 660. 

coal, 205. 

consolidated pop safety, 668. 

Crane atmospheric, 666. 

Crane hydraulic emergency, 659. 

Davis back pressure, 665. 

diaphragm, 647. 

disk, 527. 

emergency, 558. 

foot, 668. 

Foster back pressure, 665. 

gate and globe, 655. 

Illinois Eng. Company's vacuum, 643. 

Kieley reducing, 667. 

Ludlow angle, gate pattern, 657. 

non-return, 658. 

nozzle, Curtis turbine, 352. 



OCT 5 191C 



902 



INDEX 



Valves, nozzle, De Laval turbine, 332. 

nozzle, theoretical, 334. 

Paul vacuum, 645. 

reducing, 666. 

stop, 655. 

Webster vacuum, 643. 
Vanes of Curtis turbine, 352. 

Hamilton-Holzworth turbine, 362. 

Westinghouse-Parsons turbine, 365. 
Van Stone joint, 613. 
Vegetable oils, 669. 
Velocity of steam through nozzles, 328. 

through pipes, 633. 
Velocity of water through nozzles, 650. 

through pipes, 650. 
Venturi meter, 

Vertical blow-off connections, 116. 
Vertical tubular boilers, 69. 
Volatile matter in coal, 23. 
Visible smoke, loss due to, 38. 
Viscosity of oils, 673. 
Volute centrifugal pump, 561. 
Volume of jet condenser chamber, 408. 

Wainwright feed-water heater, 493. 

Wall brackets, piping, 622. 

Wanner optical pyrometer, 738. 

Warren fuel oil burner, 59. 

Washed coals, 40. 

Water and boiler scale, analyses of, 407. 

Water columns, 119. 

Water-cooling systems, 453. 

Water, flow of, in pipes, 650. 

Water, friction coefficient in clean iron 
pipes, 652. 

Water, height of lift by suction, 538. 

Water pistons, 530. 

Water-softening plants, 480. 

Water temperatures, feed-water heaters, 
490. 

Water, weight of, for condensing, 406. 

Waterworks, centrifugal pump charac- 
teristics for, 565. 

Weber concrete-steel chimneys, 235. 



Webster feed-water heaters, 488. 

Webster heating system, 406. 

Webster vacuum valve, 643. 

We-Fu-Go purifying system, 481. 

Weighing fuel, 852. 

Weighing water, 850. 

Weight of air as indicated by flue-gas 

analysis, 30. 
Weight of, 
Weight of boiler compound necessary, 

478. 
Weight of guyed steel chimneys, 219. 
Weight of water evaporated per square 

foot of heating surface, 95. 
Weighton multi-flow surface condenser, 

419. 
Weiss barometric condenser, 411. 
West Albany power station, N. Y. C. R. R., 

788-797. 
Western Electric Co.'s Power plant, 272. 
Westinghouse-Leblanc condenser, 445. 
Westinghouse-Parsons steam turbines, 

365. 
Wet air pumps, 553. 
Wheeler admiralty surface condenser, 

417. 
Wheeler, C. H., multi-flow surface con- 
denser, 443. 
White-Star oil filter, 688. 
Wickes, boiler, 84. 
Wilkinson stoker, 136. 
Williams oil burner, 58. 
Wire drawing, 254. 
Wood as fuel, 18. 
Wooley smokeless furnace, 149. 
Worthington barometric condenser, 414. 
conoidal pump, test of, 568. 
cooling tower, 458. 
jet condenser, 402. 
Wrought-iron pipe, 611. 

Yonkers power house, N.Y.C.R., piping 
for, 

Zinc, use of, in boilers, 477. 






